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Publication numberUS2093295 A
Publication typeGrant
Publication dateSep 14, 1937
Filing dateFeb 23, 1934
Priority dateFeb 23, 1934
Publication numberUS 2093295 A, US 2093295A, US-A-2093295, US2093295 A, US2093295A
InventorsWilford H Teeter
Original AssigneeWilford H Teeter
Export CitationBiBTeX, EndNote, RefMan
External Links: USPTO, USPTO Assignment, Espacenet
Compressor
US 2093295 A
Abstract  available in
Images(1)
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Claims  available in
Description  (OCR text may contain errors)

W. H. TEETER sept. 14, 1937.

COMPRESSOR original Filed Feb. 23, 19:54

ATTORNEY Patented Sept. 14, 1937 UNITED STATES PATENT OFFICE Application February 23, 1934, Serial No. 712,508 Renewed July 3, 1936 4 Claims.

This invention relates to improvements in compressors particularly adapted to be used in refrigerating systems.

It is among the objects of the present invention to provide a compressor, preferably of the two-stage type, which is adapted to provide greater volumetric and also greater overall efficiency in a refrigeration system.

A further object of the present invention is to provide a compressor so constructed and arranged as to prevent undesirable circulation of lubricating oil throughout the refrigerating system which, in most instances, is caused by the foaming of the oil due to rapid vapor pressure changes in the atmosphere contacting the lubricating oil.

A still further object of the present invention is to provide a compressor of simple structure and compact design adapted to operate eiciently with comparatively low power requirements to drive it.

Further objects and advantages of the present invention will become apparent as the description of the accompanying drawing progresses,

said drawing illustrating a preferred embodiment of one form of the invention.

In the drawing:

Fig. 1 is a longitudinal sectional view of a multi-cylinder compressor.

Fig. 2 is a transverse sectional View of the compressor, taken substantially along the line 2-2 of Fig. 1. Certain parts of the machine are shown in elevation for the sake of clearness.

Referring to the drawing, the numeral I0 designates the crankcase of the machine, which crankcase provides the cylinders II, I2 and I3. The crankcase is open at the bottom, this opening being normally closedV by the bottom pan I4, having suitable gaskets to prevent fluid leaks.

At opposite ends of the crankcase lugs I5 and I6 are provided, respectively. which lugs are bored out to provide bearings in axial alignment andin which the crank-shaft I1 is journalled. The bore in lug I5 does not extend to the outside of the crankcase and thus requires no seal, however, the bore in lug I6 extends through its entire length so as to permit one end of the crankshaft l1 to extend outside the casing I6 to permit attachment of the driving element (not shown). Any suitable seal I8 is provided between the shaft I1 and lug I6 so as to prevent fluid or gases from leaking from the crankcase at this point.

Fins I9 are provided on the cylinder portion of the casing for strengthening the structure and (ci. 23o-191) for dissipating heat generated at .these points. Cylinders II and I3 may be termed first stage compression cylinders and the one designated by the numeral I2 may be termed the second stage compression cylinder. In the present instance two iirst stage compression cylinders are shown, but it is obvious that only one is necessary to cooperate with the second stage cylinder. Each cylinder Il, I2 and I3 has a reciprocative piston 2I, 22 and 23, respectively. Each of these pistons has a port or passage in its head portion and each port is provided with a control valve. inasmuch as pistons 2I and 23 are alike, only one will be described detailedly.

The port in the head of piston 2I is designated 15 by the numeral 24. This port consists of a plurality of openings arranged about a central web 25. An annular ridge or projection 26 on the inner surface of the piston head surrounds the port 24, said ridge providing a valve-seat adapt- 20 ed to be engaged by the disc-valve 21. This valve is normally, yieldably .urged into engagement with the seat 26 by a spring 28 abutting against the head of screw 29 threaded into the web 25. The usual wrist pin is provided in the piston providing for the attachment of the connecting rod 30. For purposes of description of the operation of the device, the port of piston 23 is referred to by the numeral 3l, the valve by numeral 32 and the connecting rod of this piston is 30 designated by the numeral 33.

The piston 22, in the second stage compression cylinder I 2 has a port 35 in its head portion, similar to the ports in pistons 2| and 23. In the piston 22, however, the annular ridge 36, pro- 35 viding a valve-seat for valve 31, is on the outer surface of the piston head instead of the inner surface as in the case in pistons 2I and 23. An apertured plate 38 is secured to the outer surface of the piston head 22 in such a manner that 40 said plate is normally spaced from the valve 31 when it is in engagement with its valve-seat 36, thus permitting the movement of the valve from its seat, still preventing improper displacement of the valve from the piston. 45

The crank shaft I1 has three eccentrically arranged discs 4I,42 and 43 in the present instance arranged degrees apart, which discs form throws for the several connecting rods. A ringshaped end 46 on rod 33 ilts upon the disc 4I, a 50 similar ring-shaped end 45 on rod 39 fits upon disc 43 and the ring-shaped end 44 of rod 30 fits upon disc 42.

An oil well 50 with ducts 5I leading therefrom is provided in the lug I6 within the crank-case 55 and a similar oil well 62 with ducts 63 leading therefrom is provided in the lug I5 within the crank-case. These provide for the delivery of lubricating oil to the two bearings of the crankshaft I1.

The numeral 60 designates the cylinder head of the compressor attachable to the cylinder block or casing I0 in any suitable manner. This cylinder head provides three separate chambers 6 I, 62 and 63. Chambers 6I and 63 are in communication with the respective cylinders il and I3 while chamber 62 is in communication with the cylinder I2. Port 64, in the cylinder head block 60 connects cylinder II with chamber 6I and a similar port 65 connects cylinder I3 with chamber 63. Valves control the ow of fluid through these ports 64 and 65, and inasmuch as both valve structures are alike, only one will be described for the sake of brevity.

Referring to Fig. 1, port 64 is shown having an annular ridge 66 which provides a valve-seat for the valve 61 which is normally, yieldably urged against this seat by a spring 68 interposed between a wall of chamber 6I and an abutment collar 69 secured to the end of the stem of valve 61. For the sake of description of operation, the similar valve of cylinder I3 is designated by the numeral 10 and its spring 1I.

Both chambers 6I and 63 are connected to .a common suction line 12 through suitable nipples 13, couplings 14 and pipes 15. Chamber 62, however, is in communication with a discharge conduit 16. The port 80 which connects cylinder I2 with chamber 62 is normally closed by any suitable valve mechanism which must be opened by pressure within the cylinder I2. In the drawing this valve is shown as being a flexible blade 8| anchored at one point to the cylinder head by a screw 82. A spring 84 interposed between the blade 8| and a bracket 83 secured to the cylinder head, assist the blade 8l in normally closing the port 80.

'I'he structure of the device having been described detailedly, its operation will now be explained.

When the power element, not shown, but adapted to be attached to shaft I1, is rotating said shaft, pistons 2I, 22 and 23 are reciprocated in their respective cylinders. As the pistons 2| and 23 are moved downwardly, toward the shaft I1, the resultant displacement of the pistons creates a lower pressure in the respective cylinders than exists in the suction line 12, thereby causing the respective valves 61 and 10 to be moved from their seats against the effect of their springs 68 and 1I respectively. 'I'his results in fluid being drawn from the suction line 10 through the pipes 15, couplings 14 4and nipples 13 into the respective chambers 6I and 63 and thence through the now open ports 64 or 65 respectively. It will of course be understood that due to the angular relation of the two crankshaft throws 4I and 42 the suction cycles of the two pistons 2l and 23 are not concomitant but successive, thus uid will flow into chamber 63 and from there into cylinder I3 at a different time during a cycle of compressor operation than when said fluid flows into chamber 6I and from there through port 64 into the cylinder I I.

The first stage of compression occurs when either piston 2| or 23 moves upwardly to compress the fluid previously drawn into their respective cylinders Il and I3. The valves 61 and 16 of the said cylinders I I and I3 will close when the suction strokes of the respective pistons in said cylinders are completed. When the fluid in a cylinder is compressed to a predetermined degree the piston valve, valve 21 on piston 2 I or valve 32 on piston 23, is actuated thus permitting the compressed fluid to escape from the cylinder, through the open piston port into the crank-case in which the lubricating oil for the compressor is stored. The piston valve closes the piston ports automatically when the piston reaches its extreme upper position or upper dead center. movement of the piston repeats the suction or intake cycle thereof.

While the pistons 2l and 23 are being reciprocated in their cylinders, piston 22 is likewise being reciprocated in cylinder I2, the relative position and cycles of these pistons being controlled by the relative angular positions of their respective crank-shaft throws or eccentrlcs 4I, 42 and 43. When the second stage pressure piston 22 moves downwardly, the pressure differential on the two sides of the piston head will cause the valve 31 in said piston to be moved to open the port 35, thus permitting the fluid under the first stage of compression within the crank-case, to flow into the compressionchamber of cylinder I2. When piston 22 reaches its lowermost position, valve 31 will close port 36 in said piston. Now, as the piston starts on its upward journey it exerts pressure upon the fluid within the compression chamber of cylinder I2 which, when compression in said chamber has reached a predetermined degree, will cause valve 8| to be lifted and permit the compressed fluid in cylinder I2 to be discharged lnto the receiving chamber 62. Continued discharge of fluid under pressure into chamber 62 will force the fluid from said chamber through the discharge conduit which, in a refrigerating system, is connected to the condenser.

Pistons 2l and 23, being first stage compression pistons, will draw gases into the compressor cylinders and deliver said gases into the crank-case of the compressor under rst stage compression. Thus the crank-case may be identified as a container of first stage compressed vapor or gas. The piston 22 takes its supply from this crankcase and delivers said supply of vapor under second stage compression to a chamber, which in a refrigerating system is connected with the condenser.

The casing thus provides a crank-case in which the reciprocating mechanisms of the pistons are housed, into which the vapors are discharged under rst stage compression and from which said first stage compressed vapors are withdrawn to be compressed at the second stage for final delivery to the condenser of the refrigerating system. The crank-case is also the reservoir for the lubricant supply. Due to the fact that the lubricant supply is carried within the second stage compression chamber, its subjection to wide varying pressures and therefore its tendency t0 foam is materially reduced.

The advantages of applicants compressor may readily be recognized by comparing its design,

operation and emciency with ordinary machines of the same class. n

In the conventional refrigeration compressor, the capacity of the unit to absorb heat at low evaporator temperatures is a comparatively small fraction oi' the capacity to absorb heat at somewhat higher evaporator temperatures. For example. a unit produced by a large manufacturer and marketed in large volume has a B, t. u. capacity of approximately 1360 per hour with an evaporator temperature of 26 degrees Fahren.

Downward heit, but when the unit is operated with an evaporator temperature of -12 degrees Fahrenheit, under the same atmospheric temperature conditions, the capacity of the unit is reduced substantially to 416 B. t. u. per hour. 'I'his loss in capacity is due primarily to the fact that the unit, when operating at the colder temperature, must have a compressor suction pressure considerably below atmospheric pressure, which results in the refrigerant vapors becoming rareed, or less dense, and the weight of gas drawn into the compressor on the suction stroke of the piston is considerably less than that drawn into the compressor on the suction stroke when the refrigerant gas is dense, as is the case when operating at higher temperatures. This loss in capaoity is therefore due to loss in volumetric eiliciency of the compressor. While there are many types of refrigerants that may be used, some of which will produce low temperatures while operating under a suction pressure that is above atmospheric pressure, the general result is the same in each case; dense vapors give high volumetric efilciencies, and rareed vapors give comparatively lower volumetric eillciencies, and there is not a great deal of difference in overall eillciencies obtained on conventional'compressors due to the type of refrigerant used. 'Ihis loss in volumetric efficiency is further augmented by the wide pressure differential between the suction pressure and the cylinder head pressure. Whatever volume of vapor that remains in the cylinder after the piston has traveled to the top of its stroke, which is due to clearance around the valves, and head clearance, is compressed to maximum head pressure at the end of each compression stroke, but remaining in the cylinder, must expand on each down stroke of the piston to a value at least equivalent in pressure to the suction pressure of the compressor before any additional charge can enter the cylinder to be compressed onthe next stroke. Where the pressure differential is high between the high and low pressure side of the compressor, a greater movement of the piston is necessary to provide the necessary displacement for expansion of the gas remaining in the cylinder after the compression stroke has been completed; this movement of the piston does no useful work in drawing in a new charge, and represents a loss in both energy and displacement.

In my invention, the compressor is so designed as to insure that a very low pressure differential is always present between the low and high pressure sides of the suction, or first stage piston, in this way preventing the loss in volumetric emciency and wasted power described in the previous paragraph.

With this arrangement, it is possible to make the compressor smaller and still do the same amount of work, since a much higher percent of complete volumetric efficiency is obtained.

Another means for reducing power required to operate the compressor in my design, is the application of vapor which has been compressed to the first stage, to the under side of the second stage piston, so that the pressure differential on the two sides of the second stage piston equals the final pressure rriinus the first stage pressure.

A further reduction in the power required to 'operate my compressor results from its smaller relative size and piston displacement to do a given amount of refrigeration. With every compressor there is a certain amount of frictionl which must be overcome in order to move the parts at the operating speed, which might be termed residual friction" and the total power required to operate the compressor is that necessary to overcome the residual friction, plus the power necessary to accomplish the compression of the vapor. In small compressors used for household refrigerators and small commercial installations, the residual friction sometimes requires more power than the useful load, and is always a high percentage of the total power requirements. By reducing the size of the pistons and also reducing the bearing loads on certain parts of the compressor in the present invention the residual friction load is appreciably lessened, thus reducing power requirements, and increasing the overall eiliciency of the unit.

When compressors are used which are designed to draw the charge of vapor from the evaporator into the crankcase of the compressor and into the compression cylinder through piston valves, the compressor lubricating oil in the crankcase is subjected to rapid pressure changes when the compressor is started up after an off cycle. 'I'his causes the oil to give oil quantities of refrigerant which have been absorbed during the "off cycle when the crankcase pressure was increasing, and the oil foams violently and becomes frothy, and considerable quantities of the oil are carried through the piston valves, and must cir- .culate through the condenser, receiver and evaporator before it can return to do useful work in the crankcase of the compressor.

With some compressor designs, the amount of oil that migrates through the system is so great as to deprive the compressor of proper lubrication, and cause stuffing box seal trouble. 'This is especially true where refrigerants that are miscible in oil in any ratio, are used.

My invention provides a means of preventing the oil in the crank-case from being subjected to evaporator suction pressures, thus preventing excessive foaming of the oil and the resultant migration thereof.

It may also be noted that by the use of my invention, the variations in torque effort throughout a single revolution of the drive shaft are materially reduced due to the fact that the work cycles for each of the pistons are staggered relative to the work cycles for the other pistons. In other words, the present design considerably smooths out the torque required to drive the compressor through a single revolution since the compression of gas in the second stage cylinder does not take place simultaneously with the compression of gas in the crankcase by the two first stage pistons.

In the accompanying drawing I have shown the compressor designed as a two-stage three cylinder compressor, but it is expressly understood that my application is not limited to the details of the design as shown; that any desirable number of cylinders may be used to accomplish the action described in this application, and any number of stages of compression may be used by the use of a web in the crank-case separating the stages used in excess of two, and sealed from the other crank-case compartments by means of stuffing box seals.

While the embodiment of the present invention as herein disclosed, constitutes a preferred form, it is to be understood that other forms might be adopted, all coming within thescope of the claims which follow.

What is claimed is as follows:

1. A multi-stage compressor comprising a first stage compression section including at least one cylinder having an inlet check valve in the head thereof and an outlet, a second stage compression section including a cylinder having an inlet and an outlet, a crankcase communicating with the outlet from the rst stage cylinder and with the inlet to the second stage cylinder, a drive shaft, means in said ilrst stage cylinder for intermittently compressing uid, means in the second stage cylinder for intermittently compressing fluid and means including separate connections between each of the compressing means and the drive shaft for staggering the work cycle of each compressing means with respect to each other compressing means.

2. A multi-stage compressor comprising a rst stage compression section including at least one cylinder having an inlet check valve in the head thereof, a movable piston in said rst stage cylinder, said piston having an opening therein constituting an outlet from said cylinder, a check valve for controlling said outlet, a second stage compression section including a cylinder having an outlet check valve in the head thereof, a piston in said second stagefcylinder, said last named piston having an opening therein constituting an inlet to said second stage cylinder, a checkl valve for controlling said inlet to said second stage cylinder, a crankcase communicating with the outlet from said first stage cylinder and with the inlet to said second stage cylinder, a drive shaft, and means including separate connections between each of the pistons and the drive shaft for staggering the work cycle of each piston with respect to the work cycle of each other piston.

3. A multi-stage compressor comprising a pluralityl of first stage compression chambers each having a valve controlled inlet and a valve controlled outlet, a second stage compression chamber having a vvalve controlled inlet and a valve controlled outlet, said compressor having a crankcase communicating with all the outlets of said first stage compression chambers and with'the inlet to saidsecond stage compression chamber. separate movable compressing means in each of said chambers. a drive shaft,.and means including separate connections between each `of the compressing means and the drive shaft for staggering the work cycle of each compressing means with respect to each other compressing means.

4. A multi-stage compressor comprising a iirst stage compression section including at least two cylinders each having an inlet check valve in its head, movable pistons in said cylinders, each piston having an opening therein constituting outlets from said cylinders, a check valve for each of said openings, a second stage compression section including a cylinder having an outlet check valve in the head thereof, a piston in said cylinder, said last named piston having an opening constituting an inlet to said second stage cylinder, a check valve for controlling said inlet to said second stage cylinder, a crankcase communicating with the outlets from all of said ilrst stage cylinders and with the inlet to said second stage cylinder, a drive shaft, and means including separate connections from eachof said pistons to said drive shaft for staggering the work cycle of each piston with respect to the work cyclel of each other piston.

WILFORD H.- mm.

Referenced by
Citing PatentFiling datePublication dateApplicantTitle
US3204865 *Feb 6, 1964Sep 7, 1965Theodore Neubauer EmilTwo-stage refrigerant compressor
US4542768 *Mar 12, 1984Sep 24, 1985Rotron, Inc.Pressure relief valve
US6776587 *Dec 20, 2000Aug 17, 2004Knorr-Bremse Systeme für Schienenfahrzeuge GmbHDual-stage, plunger-type piston compressor with minimal vibration
US7645125Jun 22, 2004Jan 12, 2010Delphi Technologies, Inc.Refrigerant compressor with improved oil retention
US20130078122 *Nov 16, 2012Mar 28, 2013Ams R&D SasDiaphragm Circulator
Classifications
U.S. Classification417/248, 417/255, 137/857
International ClassificationF25B31/00
Cooperative ClassificationF25B31/00
European ClassificationF25B31/00