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Publication numberUS3265293 A
Publication typeGrant
Publication dateAug 9, 1966
Filing dateJan 10, 1966
Priority dateSep 8, 1959
Also published asDE1403596A1
Publication numberUS 3265293 A, US 3265293A, US-A-3265293, US3265293 A, US3265293A
InventorsBenedictus Schibbye Lauritz
Original AssigneeSvenska Rotor Maskiner Ab
Export CitationBiBTeX, EndNote, RefMan
External Links: USPTO, USPTO Assignment, Espacenet
Vacuum pump of the screw rotor type and method for operating the same
US 3265293 A
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Description  (OCR text may contain errors)

llg 9, 1966 l.. B. scHlBBYE 3,265,293

VACUUM PUMP OF THE SCREW ROTOR TYPE AND METHOD FOR OPERATING THE SAME Original Filed Aug. 29, 1960 United States Patent O VACUUM PUMP F THE SCREW ROTR TYPE 1s claims. (ci. 23o-14s) This application is a continuation application replacing my co-pending application Serial No. 320,568, iiled October 31, 1963, as a continuation of my prior application Serial No. 52,485, filed August 29, 1960, both now abandoned, land relates back to said applications for all dates and rights incident to the filing thereof land of the corresponding Swedish application Serial No. 8,297/59, iiled September 8, 1959, from which priority is claimed.

The present invention relates to vacuum pumps and more particularly to vacuum pumps of the screw rotor type, and has for its general objects the provisions of novel and improved methods and apparatus for the creation and maintenance of sub-atmospheric pressures in pressure vessels or systems requiring such pressures, such, for example but without limitation, as condensers utilized in steam turbine power plants and reaction vessels of various kinds utilized in chemical processes involving steps required to be effected at sub-atmospheric pressures.

The eld with which the invention is particularly concerned and to which it is primarily directed is that in which the vacuums are within the range practically feasible of attainment with reciprocating mechanical vacuum pumps even though the invention has -a wider range extending down to vacuums corresponding to somewhat lower absolute pressures. The lower limit of the range of the reciprocating pumps is represented by vacuums corresponding to absolute pressures of some 71/2 to 10 millimeters of mercury (see Marks Mechanical Engineers Handbook, Fifth edition, 1951, page 1890), which values -are representative of absolute pressures substantially lower than those required for the majority of commercial installations such, for exlaimple, as steam condensers for turbine plants and the like, the most emcient operating pressures for which may be of the order of one or more inches of mercury (see Marks Handbook noted above, page 1173, giving a table of such values for the most economical vacuum vfor the majority of installations).

In steam condensing plants and the like, vacuum pumps of relatively large capacity are required `for the evacuation of the relatively large vessels required, in order to bring the pressures down to the desired level, and this large capacity must be coupled with relatively high efficiency of operation over long periods lat the normal subatmospheric operating pressure when the actual mass of fluid being evacuated is relatively very small.

Because of their relatively high speed of operation and high volumetric capacity for a given size off compressor, it was thought that compressors of the helical screw rotor type, more commonly known as screw compressors, would be highly useful as vacuum pumps for producing and maintaining vacuums within the range under consideration, but for reasons hereinafter more fully appearing, it was found that such compressors in their most simple and economical form, that is, single stage compressors, or when designed in the form of conventional two-stage compressors, would not economically be feasible or satisfactory.

It is accordingly the general object of the present in- Patented August 9, 1966 ICC vention to provide improved screw rotor compressor apparatus and improved methods of oper-ating the same which make the use of such apparatus not only highly ellicient, but also desirable from 4the standpoint of size, weight and cost vfor a given capacity as compared with the comparable equipment of other types presently available for the purpose.

In order more fully to understand the nature of the invention and the manner in which its objects are accomplished, it is first desirable to review briefly some of the fundamental principles of operation of the screw compressor and the reasons why a single-stage compressor of this type is not suitable for use .as -a vacuum pump for creating and maintaining Vacuums of the order under consideration.

For this purpose, a most appropriate example of such apparatus is that disclosed in Nilsson Patent No. 2,622,- 787, granted December 23, 1952, a modied version of the construction of which constitutes the embodiment of apparatus herein disclosed by way of example for carrying the present invention into effect. The known conventional form of this apparatus comprises a housing structure, enclosing two intermeshing rotors having helical lands and grooves forming between themselves and the housing structure chevron-shaped compression chambers decreasing in volume between the time of cut-off from an inlet port in the housing to the time of communication with a delivery port in the housing spaced from the inlet port. It is basically characteristic of these apparatus that a running :clearance is maintained between the circumferences of the rotors and the bores in the housing structure in which they operate and also that in many instances the rotors are turned in timed relation with respect to each other through the medium of timing gears which maintain a running clearance between the intermeshing lands and grooves of the rotors. In some instances the timing gears may be dispensed with, but in all cases the basic characteristic which is most important to the present invention is that the compression chambers are apositively sealed. With apositive sealing, in contrast with positive sealing, such `as that effected by the piston rings of a positively sealed Ireciprocating piston compressor, volumetric eciency of the machine is dependent upon the extent of the pressure rise in the compression chambers during any one cycle, or in other words, the value of the compression ratio, since the leakage from the apositively sea-led compression chambers will obviously increase with increase in the pressure rise in a single stage.

If we now consider the lower limits of the range of vacuums with which we are concerned, it is evident that if the vacuum is to be carried down to some 71/2 to l0 millimeters of mercury, compression of the evacuated fluid through a compression ratio of from 75-100 to 1 is required in order to provide discharge at atmospheric pressure. If we now consider the utilization of -a single-stage screw compressor for effecting this compression ratio, experience has shown that with normal clearances the volumetric efficiency of such compression would not exceed some 20%, with there being some leakage from the Iapositively sealed compression chambers. This in itself would call for a compressor of inordinately large size for a given capacity, but what is more important, the very small variations within the dimensional tolerances that must be allowed for in any commercially feasible manufacture, have an inadmissibly large negative effect on such a machine. If, for example, specified tolerance variation permits a variation in the clear-ance spaces of only 10%, which from a manufacturing standpoint is about :as low a Variation as is commercially practical, a 10% increase in the clearances will, under the assumed condition increase the leakage from 80 to 88%, leaving a net volumetric erliciency of 12%. Thus, if we assume `a normal efficiency of 20% and must allow for a 10% manufacturing variation in the clearances, we must increase the designed size of the compressor by 67%, in `order to guarantee rated capacity from a compressor with only 10% larger than normal clearances, for which allowance must be made. From the Vforegoing it is thus evident that out from the stand-point of eiciency and from the standpoint of the weight and size of the equipment required for a given capacity, a single-stage screw compressor is not a competitively advantageous form of apparatus for use as a vacuum pump for vacuums of the order under discussion.

However, the present invention provides `a very simple and effective solution to this situation by providing a new and improved form of two-stage compressor of the screw rotor type, which effectively overcomes the deficiencies of the single-stage type and which enables such Ia reduction in the size of the compressor to be effected, in relation to the size of a single-stage compressor of the same capacity, that the cost of a two-stage compressor of a given size may be not greater and can be even less than lthat of a single-stage compressor having the capacity for the same duty. In accordance with the broad concept of the invention, the total pressure rise required to deliver against atmospheric pressure from a desired inlet pressure, is divided into two stages, the second stage of which may `be termed the main compression stage, since this stage will operate through much the higher compression ratio, and which m-ain stage is preceded by what may be termed a pre-compression stage, having a relatively very low compression ratio, and consequently, a relatively very high volumetric eiciency, and delivering the pre-compressed fluid at a relatively very low intermediate pressure to the second compression stage. Specifically, the compression ratio for the initial or pre-compression stage, should not eX- ceed to 1 and preferably should be as low as of the order of 2 or 3 to 1. If now we consider for the moment a pre-compression stage in which the pressure ratio is only 2 to 1, the volumetric efficiency of a normal compressor will be of the order of 95%. Now, if the same manufacturing variation in clearances of is assumed, it is evident that this variation will affect the volumetric efficiency and capacity of the pre-compression stage to the extent of only 0.5%. The pressure ratio of the second or main stage of such a twostage compressor will operate over only half that of a singlestage compressor for the same purpose. As the internal leakage in a screw compressor is about proportional to the pressure ratio therein the leakage will go down from 80% for the conditions mentioned above to about 40%, which means that the volumetric efficiency goes up from to 60%. The swept Volume of the main stage of such a two-stage compressor will thus as the volume of the working uid in the inlet of the main stage will be only about half that of the precompression stage, if for the sake of simplicity the temperature rise due to compression is disregarded, be about through the compressor is assumed to be constant the intermediate pressure between the two stages increases so that the pressure ratio of the preacompression stage increases from 2 to 1 to 2.2 to 1 resulting in a decrease of the volumetric eiciency of the pre-compression stage from to 94.5%. The net variation of 'the volumetric efciency in the pre-compression stage due to 10% manufacturing variation in clearance will thus be only about 1%. The capacity of the two-stage compressor will thus vary with only 1% compared with the 40% variation of the corresponding single-stage compressor. As mentioned above the rotor length of the main stage will be only about 40% of the designed length of the single stage rotor length so that the total rotor length of the pre-compression `and main stages including intermediate shaft portions will not exceed the rotor length of the single-stage compressor at the same time as the rotor diameter of the two-stage compressor will be only about 40% of that of the single-stage compressor. The overall dimensions of the two-stage compressor will thus be considerably smaller than those of ithe single-stage compressor and consequently the costs of the two-stage compressor will be lower than those of the single-stage compressor.

Looked at in another way, the pre-compression stage can be considered as a highly efficient non-return valve delivering a much reduced volume of evacuated fluid to the main compression stage for subsequent compression through a much higher compression ratio at -rnuch lower efficiency, the latter, however, not being of material significance since the highly efficient pre-compression stage is determinative of the volumetric capacity of the apparatus and consequently the size required for a given duty. The invention further contemplates in its preferred form the use of an apostively sealed pre-compression stage of the screw rotor type and further as a preferred embodiment of that form contemplates the employment of a pre-compression stage, the rotors for which are in tandem with the rotors of the m-ain stage, with the rotors of like -diameter and helix angle for obvious reasons of manufacturing facility and low cost.

With the rotors arranged in tandem and connected in series and with the rotor diameters and profiles in the two stages being the same, the displacement or swept volumes of the two stages are proportional to the lengths of the rotors in the respective stages. Since the stages are connected in series, the mass ow through the different stages must `be eX-actly the same so that to obtain the value of the real volume of the fluid handled, these swept volumes must be corrected by relating them to the volumetric eflciencies of the respective stages.

There is to be noted that comparatively very small variations in the value of the a-bsolute inlet pressures result in comparatively large variations in the compression ratios required to elevate the inlet pressure up to atmospheric discharge pressure, and insofar as the swept volumes of the respective stages are concerned, and their relation to each other, it can be shown that with decreasing values of the inlet pressure increasing swept volume of the main compression stage in relation to that of the precompression stage is required. In other words, with a design of the kind presently assumed, the lower the inlet pressure for which the apparatus is designed, the greater must be the length of the rotors of the main compression stage relative to the length of the rotors of the pre-compression stage.

The foregoing relationship may be demonstrated by the `following calculations, which for the sake of simplicity rdisregard temperature rise due to compression, which simplification while affecting specific values, does not alter the nature of the variation in the relationship between the relative rotor lengths of the two stages.

As the two stages are designed in such a way that the rotor diameter and the rotor profile in the two stages are exactly the same the swept volumes of the two stages may in a comparison therebetween be represented only by the length of the rotor in each stage, LI and LH, spectively. The real volume of the passing fluid, however, has to be determined by correction with regard to the volumetric eiciency of each stage, avon and mom, respectively. The mass of fluid passing each stage will be represented by the product of the real volume of the fluid and the inlet pressure to said stage, the inlet pressure p1 and the intermediate pressure pmt, respectively. Since as said above the mass ow through the two stages must be the same the following eq-uation lmust be true WvoiyLp1=77vol11L11pint LI mi LII'UvolII p1'7vo1I However, if it is assumed that the pressure ratio of the rst stage,

Pint

is constant also the volumetric efliciency of this stage, oM011, is constant. Thus u const p1'7lvol1 and consequently L 1- const LII'VOIII If two different two-stage compressors designed to have the same compression ratio in the pre-compression stage are designed to provide different over-all compression ratios, the one designed for the greater over-all compression ratio must effect in the main stage a greater compression ratio at a lower volumetric eciency than is the case for the main stage of the other compressor and thus in accordance with the foregoing formula requires a lower value `for the ratio Ll/Ln, or, in other words, the higher vacuum for which the pump is designed the greater should be the length of the rotors of the main stage in relation to the length of the rotors of the precompression stage.

It is further to be noted in connection with the matter of swept volumes and compression ratios, that the apositive sealin-g characteristic of the screw compressor, provides automatic adjustment of the actual compression ratio being effected in the two stages relative to each other and relative to the over-all compression ratio required to effect the necessary pressure rise in order to deliver against a substantially constant delivery pressure from an inlet pressure which may be variable. In any compressor design a certain xed in-built compression ratio is provided, this ratio Vbeing determined by the decrease in the volumes of the compression chambers between the point of cutoff from the inlet port and the point of communication with the discharge port. In case the actual inlet pressure (assuming a constant delivery pressure) varies materially from that for which the in-built compression ratios were designed, the apositive nature of the sealing results in such variation of the actual rate of leakage from the chambers as to produce automatically whatever actual' FIG. l is a longitudinal section, partly in elevation, of a compressor embodying the invention and designed for creating and maintaining a vacuum corresponding to an absolute pressure of 71/2 to l0 millimeters of mercury and,

FIG. 2 is a section taken on line 2 2 of FIG. 1.

Referring now more particularly to FIG. l of the drawings, the housing structure indicated generally at 10, comprises a barrel portion 12 separated into low pressure and high pressure sections 12a and 12b, respectively. The bores of the low and high pressure sections of these, of which the low pressure sections 14a and 16a are seen in IFIG. 2, are in alignment.

The barrel portion of the housing structure is closed at the losw pressure end of the compressor by nie-ans of the end closure member 18 providing the inlet 20 for the admission of the Huid to be compressed. The high pressure end of the barrel portion is closed by a high pressure end member 22 providing the discharge or outlet 24.

Rotatively mounted in the end members by suitable bearings are the male rotor 26 and the -femlale rotor 28, each of these rotor-s being divided into low pressure and high pressure sections of which the iow and high pressure sections of the male rotors 26a and 26h, are seen in FIG. 1.

The low pressure section of the housing structure is provided with la discharge port 30, the high pressure section being provided with an inlet port 32, and these ports |being appropriately connected for series flow of the working fluid through the apparatus by means of the conduit 34.

The rotors 26 and 28, of ythe male and female type char-acteristic of such compressors, may be of any suitable conguration or profile appropriate for compressors of this kind, of which several specific profiles are well known in the art and of which the proiiles shown in the aforementioned Nilsson Patent 2,622,787 is one, and which is here .illustrated in FIG. 2 .by ,way of example. The low and high pressure sections of the rotors, such as sections 26a and 26b shown in FIG. l, are separated by an intermediate rotor shaft part passing through a central diaphragm or web in the housing structure vand in which is located a suitable seal of any lappropriate known design -for preventing the passage of flow of fluid from one to the other of lthe sect-ions -of the compressor. As an aid to insuring a tight .seal 36 at this place, a sealing liquid may be introduced as is more or less diagrammatically indicated by the feed pipe 3S yfor liquid.

In the present instance, close running clearances are provided in known conventional fashion :between the inter-meshing rotors, with or without the aid of timing gears such as shown at 40 in FIG. l, and between the peripheries and ends of the rotors Iand the enclosing housing structure. All of this is old `and well known in the art and therefore need not be further described in detail herein.

Also, as is oid in the art, liquid -is advantageously introduced into the working space of the compressor for the purpose of aiding in sealing the running clearance spaces and also for directly cooling the contents of the compression chambers to reduce the temperature rise thereof as the work of compression is done thereon. By way of illustration the introduction of such -liquid is indicated diagrammatically by the supply pipe 42 delivering a spray of liquid into the compressor intake, although it wil-l be understood that other and equivalent known means and methods for introducing liquid into lthe compressor, -such for example, as that disclosed in Nilsson et al. Patent 3,129,877 may be employed.

If desired, liquid deriving from condensated vapour may be carried oft' from -the conduit 34 interconnecting the two compressor stages, for instance, as diagrammatically indicated in FIG. l, to a steam condenser 44 from which the fiuid to be compressed may be derived.

For sealing and cooling .as well as condensing purposes, any appropriate liquid may be employed depending upon the nature of the elastic iluid being evacuated, and the fluid, if any, entrained therewith. In the case of a condenser for a steam turbine plant, water -is of course appropriate, whereas for example in ythe case of the vacuum casting of steel, lubricating oil may be indicated. `In some cases, as for instance, in the chemical industry, neither oil nor water may be suitable, and in such instances other liquids such for example as products of `apiezones or the like, may be employe-d. In other words, the speci-fic composition of the liquid is not germane to the invention.

In the illustrated embodiment, the inbuilt compression ratio for the pre-compression or low pressure section of the compressor is in the preferred range of 2 or 3 to 1 and for rthe designed inlet pressure above noted the high pressure or main compression section is provided with rotors 1.2 times the length of the rotors of the loiw pressure section, the low pressure section having rotors, the lands and grooves of which have a tot-al wrap angle of 250 in the male rotor, while the total wrap angle of the lands and grooves of the male rotor of the high pressure section is 300. Other things being equal, if the design were intended for a higher inlet pressure, .such for example, as a pressure of one inch of mercury which is lan appropriate pressure .as the inlet pressure `from a steam condenser, the length of the rotors of the high pressure section would be less relative to that of the rotors of the low press-ure section and might well result in `a compressor designed in which the high pressure section is .actually shorter rather than longer than the low p-ressure section.

As previously noted, the pressure ratio for the precompression stage should not exceed 5 to 1 and is preferably considered below that since the benefits to be derived from the high volumetric efficiency of a low compression ratio are rapidly diminished as the compression ratio is raised. Likewise, as has previously ibeen noted, the total pressure rise between inlet and delivery pressure rather than compression ratio is the most useful value to be considered in determining what may be said to be the utility yrange of the invention, since small variations in the actual val-ues of the inlet pressures result in very la-rge variations in `the values of the overall compression ratios required to deliver .to atmosphere. Accordingly, it may be said that for the purposes of obtaining to an adequate degree ythe benefits to be derived from the concept of the present invention, not only should the design and operation of the apparatus be carried out so that not over a 5 to 1 compression ratio is embodied in the precompression stage of the compressor, but the design and operation should y also be related to -the total pressure rise so that the pressure rise effected in the pre-compression stage of the compressor does not exceed of the total pressure rise, which represents approximately the percentage that would be obtained if a pre-compression ratio of as high as 5 to 1 were employed in an installation where the vacuum pump maintained an inlet pressure of one inch of mercury.

If the pressure ratio maintained in the precompression stage is in the preferred range of 2 or 3 to 1, the percentage of the total pressure rise that is effected i-n the pre-compression stage will be materially below the 15% limit, even if the installation is operated with an inlet pressure materially above one inch of mercury, which may be, according to the table previously referred to, the most economical pressure for a steam turbine condenser operating with condensing water of higher than usual temperature.

Preferably, however, the design and operation of the apparatus should be carried out so that the pressure ratio of the pre-compression stage will be within the range of 2 or 3 to 1 and so that the pressure rise effected in the pre-compression stage will not exceed the order of 5% of the total pressure rise, representing approximately the range between 3.5% and 7% that would be obtained if the pre-compression ratio was 2 to 1 and 3 to l, respectively, and the pressure i-n the inlet of the Vacuum pump was one inch of mercury.

For the pressure inthe inlet of the vacuum pump within the mentioned range of 71/2 to 10 millimeters of mercury and for the pressure ratio of the pre-compression stage within the range of 2 or 3 to 1 the pressure rise in the precompression stage will be within the range of 1% to 3%, or about 2%, of the total pressure rise which percentage is the preferred one for this range of vacuum.

While two stage screw compressors with rotor sections arranged in tandem on common shafts have heretofore been suggested and employed, the designs for such compressors have invariably been such as to produce substantially equality in the amount of work done by the different stages, with the low and high pressure stages each operating with substantially the same compression ratio. Such compressors, operating with substantially equal icompression ratios in the different stages ordinarily obtain their desired characteristics through differences in design of the two rotor sections, such differences in some cases Ibeing differences in rotor diameters and in other cases, being differences in helix angles, and total wrap angles or both. In all such prior cases, however, rthe designs are radically different from that resulting from the concept of the present invention, with the percentage of the total pressure rise effected in the low pressure stage being very much greater than contemplated by this invention.

It will be evident from the foregoing that, depending upon the nature of the uids being evacuated and the desired value of the normal inlet pressure to be maintained, there may be within the scope of the invention a substantial Variation in the value of the compression ratio to be employed in the pre-compression stage and in the percentage of the total pressure rise represented by the pressure rise effected in the pre-compression stage. Also, it will be evident that while from the standpoint of obtaini-ng the maximum advantage from the principles of the invention, two-stage compressors with each stage of the apositively sealed type provided by tandem rotors of the same diameters and lead angles are to be preferred, other specific forms and designs of apparatus may be employed to implement the principles of the invention, which accordingly is to be understood as embracing all forms of apparatus and modes of operation thereof falling within the scope of the appended claims.

I claim:

1. A multiple stage vacuum pump for compressing an elastic fluid from a sub-atmospheric inlet pressure to a delivery pressure for discharge to ambient atmosphere, comprising a main compression stage of the helical screw rotor type having intermeshing male and female rotors providing apositive sealing and further comprising a precompression stage for pre-compressing said fiuid from said inlet pressure and delivering it at an intermediate pressure to said main compression stage, said pre-compression stage having a pressure ratio not exceeding 5 to 1 and operating to effect a pressure rise in said pre-compression stage not in excess of 15% 0f the total pressure rise from said inlet pressure to said delivery pressure.

2. A vacuum pump as defined in claim 1, in which said `pre-compression stage comprises an apositively sealed stage of the screw rotor type having intermeshing male and female rotors.

3. A vacuum pump as defined in claim 2, in which said pressure rise effected in the pre-compressing stage is not lin excess of the order of 5% of the total pressure rise.

4. A vacuum pump as defined in claim 3, in which said pressure rise effected in the pre-compression stage is about 2% of the total pressure rise.

5. A vacuum pump as defined in `claim 2, in which the rotors of the two stages comprise different rotor sections mounted in tandem on two rotor shafts,

6. A vacuum pump as defined in claim S, in which a substantially positive seal is provided between the two rotor sections of each of the rotor shafts.

7. A vacuum pump as defined in claim 6, in which said positive seal is a shaft seal and a liquid is introduced as a sealing agent for said shaft seal.

8. A vacuum pump as defined in claim 5, in which the two rotor sections of each of said rotor shafts have the same diameter and profile.

9. A vacuum pump as defined in claim 8, in which the two rotor sections of each of the rotor shafts have the same helix angle.

10. A vacuum pum-p as defined in claim 7, in which liquid is supplied to the compression chambers of the pump as a sealing agent.

11. A vacuum pump as delined in claim 3, in which the compression ratio of the pre-compression stage is of the order of 2 or 3 to 1.

12. A method of evacuating elastic uid from a subatmospheric inlet pressure to a delivery pressure for discharge to ambient atmosphere, by the aid of a multiplestage apositively sealed screw rotor compressor apparatus having a main compression stage with intermeshing male and female rotors, which comprises pre-compressing the uid in an initial precompression stage through a pressure ratio not exceeding 5 to 1 from said inlet pressure to an intermediate pressure and delivering said fluid at said intermediate pressure to said main compression stage in which the fluid is compressed from said intermediate pressure to said delivery pressure, the pressure rise from said inlet pressure to said intermediate pressure effected in said pre-compression stage not exceeding 15% of the total pressure rise from said inlet pressure to said delivery pressure.

13. The method as defined in claim 12, in which the pressure rise eiected in said pre-'compression stage is not in excess of the order of 5% of the total pressure rise.

14. The method as delined in `claim 13, in which the pressure rise effected in said pre-compression stage is about 2% of the total pressure rise.

15. A method as defined in claim 13, in which the pressure ratio through which said iiuid is precompressed to said intermediate pressure is of the order of 2 or 3 to 1.

References Cited bythe Examiner UNITED STATES PATENTS 1,698,802 1/ 1929 Montelius 103-128 1,734,779 11/1929 Randolph 103-126 2,492,075 12/ 1949 Van Atta 230--158 2,592,476 4/1952 Sennet 103-128 2,622,787 12/ 1952 Nilsson 230-143 2,691,482 10/1954 Ungar 230-143 2,693,763 11/ 1954 Sennet 103-128 2,721,694 10/1955 Van Atta 230-158 2,975,963 3/ 1961 Nilsson 230--143 3,073,514 1/1963 Bailey et al. 230-143 3,074,624 1/ 1963 Nilsson et al 230-143 3,084,851 4/1963 Schibbye et al 230-143 3,121,530 2/1964 Lorenz 230-158 3,129,877 4/ 1964 Nilsson et al 230-143 3,138,320 6/1964 Schibbye 230-143 FOREIGN PATENTS 832,386 4/ 1960 Great Britain.

MARK NEWMAN, Primary Examiner.

W. I. GOODLIN, Assistant Examiner.

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Referenced by
Citing PatentFiling datePublication dateApplicantTitle
US3462072 *May 2, 1968Aug 19, 1969Svenska Rotor Maskiner AbScrew rotor machine
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US5641280 *Dec 8, 1993Jun 24, 1997Svenska Rotor Maskiner AbFor oil-free air
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US6991440 *Sep 30, 2003Jan 31, 2006Ghh-Rand Schraubenkompressoren GmbhTwo-stage screw compressor
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US7452191Apr 29, 2003Nov 18, 2008Piab AbVacuum pump and method for generating sub-pressure
US7722346 *Dec 27, 2007May 25, 2010Mayekawa Mfg. Co., Ltd.Oil supply method of two-stage screw compressor, two-stage screw compressor applying the method, and method of operating refrigerating machine having the compressor
US8257456Jan 29, 2008Sep 4, 2012Korea Pneumatic System Co., Ltd.Vacuum system using a filter cartridge
US8277207Dec 17, 2009Oct 2, 2012Mayekawa Mfg. Co., Ltd.Oil supply method of two-stage screw compressor, two-stage screw compressor applying the method, and method of operating refrigerating machine having the compressor
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DE19543879C2 *Nov 24, 1995Feb 28, 2002Guenter KirstenSchraubenverdichter mit Flüssigkeitseinspritzung
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WO2003093678A1 *Apr 29, 2003Nov 13, 2003Piab AbVacuum pump and method for generating sub-pressure
Classifications
U.S. Classification418/1, 418/9, 418/100
International ClassificationF04C27/00, F04C23/00, F04C29/00, F04C18/16
Cooperative ClassificationF04C23/001, F04C27/009, F04C29/0007, F04C18/16
European ClassificationF04C23/00B, F04C29/00B, F04C27/00E2, F04C18/16