US3927890A

US3927890A - Rotating element fluid seal for centrifugal compressor

Info

Publication number: US3927890A
Application number: US398352A
Authority: US
Inventors: Jr Maurice L Adams
Original assignee: Westinghouse Electric Corp
Current assignee: Daikin Applied Americas Inc
Priority date: 1973-09-18
Filing date: 1973-09-18
Publication date: 1975-12-23
Anticipated expiration: 1992-12-23
Also published as: US4132416A

Abstract

A dynamic seal between the high pressure discharge chamber of a centrifugal gas compressor and a sump chamber in which the driving machinery for the impeller of the compressor is contained, the sump chamber in normal operation being maintained at substantially the same pressure as the pressure at the inlet of the compressor impeller, the seal being formed by opposing faces of a seal stator and a seal rotor, and being vented at an intermediate location along the axial extent of the seal inwardly through the rotor and to the space at the inlet or suction side of the impeller.

Description

U United States Patent 11 1 1111 3,927,890

Adams, Jr. Dec. 23, 1975 [5 ROTATING ELEMENT FLUID SEAL FOR 3.004.782 l/l96l Meermans 277/1310. 11 CE T COMPRESSOR 3,392,910 7/1968 Tanzberger 415/1 1 1 3,635.57) l/l972 Wood 4l5/l22 [75] Inventor: Maurice L. Adams, Jr., Pittsburgh,

Primary ExaminerSamuel B. Rothberg [73] Assignee: Westinghouse Electric Corporation, Alwmey, g 0r m C ren Pittsburgh, Pa. 22 Filed: Sept. 18, 1973 [57] ABSTRACT I A dynamic seal between the high pressure discharge Appl' 398352 chamber of a centrifugal gas compressor and a sump chamber in which the driving machinery for the impel- 52 US. 01. 277/29; 277/1310. 23; 277/53; ler of the compressor is contained. the p Chamber 277/70; 415/111 in normal operation being maintained at substantially lnt. Cl. F16J 9/00 the Same Pressure the Pressure at the inlet of the 58 Field of Search 277/19, 18, 29, 53, DIG. 8, compressor impeller. the seal being formed y PP 277/ 32 12 415 111 ing faces of a seal stator and a seal rotor, and being vented at an intermediate location along the axial ex [56] References Ci d tent of the seal inwardly through the rotor and to the UNITED STATES PATENTS space at the inlet or suction side of the impeller. 2,938,661 5/l960 Ricketts 415/111 7 Claims, 3 Drawing g s 40 11 35 L as 52 36 OlL 1 1 i i l o o s 24 I2 H.141 5| US. Patent Dec. 23, 1975 Sheet 1 of3 3,927,890

Fm mOFE US. Patent Dec. 23, 1975 Sheet 2 of3 3,927,890

mdl

ROTATING ELEMENT FLUID SEAL FOR CENTRIFUGAL COMPRESSOR CROSS REFERENCE TO RELATED APPLICATION A seal arrangement which in some respects and in certain environments is considered to provide improved performance relative to a seal embodying the basic concept of this patent application is claimed in Adams and Raimondi U.S. Pat. application Ser. No. 398,35 1, filed concurrently herewith.

BACKGROUND OF THE INVENTION 1. Field of the Invention This invention pertains to the art of rotary or dynamic seals particularly useful in the field of centrifugal refrigerant gas compressors.

2. Description of the Prior Art U.S. patents of which applicant is aware and which have some similarities with respect to seals and/or venting passages are: U.S. Pat. No. 838,744; U.S. Pat. No. 3,392,910; U.S. Pat. No. 3,480,284; U.S. Pat. No. 3,493,169; U.S. Pat. No. 3,647,313. It is applicants view that the teachings in these patents do not anticipate or make obvious the invention claimed herein.

From the standpoint of structural similarities of certain of the parts, the closest prior art which applicant is aware is the centrifugal refrigerant compressor units currently produced by applicants assignee. These devices provide an environment in which applicants invention finds one of its most useful applications. The general arrangement of the centrifugal compressor impeller and shaft assembly and the assembly including the driving motor, sump chamber in which the drive train connecting the motor to the impeller is located and the general casing arrangement, are shown in U.S. Pat. Nos. 3,601,501 and 3,619,086, to which reference should be had for a full understanding of one general construction of a centrifugal refrigerant compressor in which the present invention is incorporated.

The present seal between the high pressure discharge space and the sump chamber on the opposite side of the wall separating these spaces is provided by a seal rotor and seal stator providing an ordinary serrated bushing type seal. The leakage of refrigerant gas from the high pressure discharge chamber side through such a seal to the sump chamber, which is maintained at substantially the same pressure as the suction pressure to the compressor inlet by venting the sump chamber to the inlet side of the impeller through an oil-refrigerant separator and oil filter arrangement, is sufficiently high with the relatively small volume sump chamber used on the relatively compact centrifugal compressor assembly, that periodically the system must be shut down to permit oil which has accumulated in the oil filter accumulator to be returned to the sump chamber. The excessive rate of refrigerant gas leakage from the high pressure side to the sump chamber creates the problem which this invention is intended to solve. A further explanation of the arrangement as a whole and the disadvantages accruing from the improved seal arrangement according to this invention will be set forth hereinafter.

SUMMARY OF THE INVENTION In accordance with the invention, the seal arrangement is provided in the described environment by a seal rotor arranged to rotate with the impeller shaft and an annular seal stator around the rotor, with opposing faces of the rotor and stator forming a dynamic seal having one end of the axial extent of the seal exposed to the compressor discharge pressure space and the other end of the axial extent of the seal exposed to the sump chamber space, and means are provided to define a venting passage from an intermediate location along the axial extent of the seal, with the passage extending inwardly into the rotor and to the inlet side of the impeller, with the venting passage having a lower resistance to fluid flow between the intermediate location along the axial extent of the seal and either end of the seal.

DRAWING DESCRIPTION FIG. 1 is a longitudinal sectional view of a prior art centrifugal gas compressor assembly of one type in which the invention may be incorporated;

FIG. 2 is a generally schematic view of the parts of a refrigerant system in which the invention is particularly applicable; and

FIG. 3 is a sectional view, enlarged relative to the showing of FIG. 1, of one form of seal arrangement as incorporated in the impeller, shaft, and casing construction according to the invention.

DESCRIPTION OF THE PREFERRED EMBODIMENTS The centrifugal compressor assembly illustrated in FIG. 1 is of a design in which compactness is considered an important feature. This is achieved in part by the use of a relatively small, but high speed, compressor arrangement. The compressor impeller in the particular arrangement illustrated is driven currently at a nominal speed of 34,000 rpm. After a brief description of the main parts of the compressor assembly of FIG. 1, the significance of the compactness relative to the invention will be discussed.

The illustrated assembly includes a compressed gas collecting scroll 10 with a shrouded centrifugal impeller ll concentric therewith secured to the end of a rotatable shaft 12 rotated by its pinion gear 13 which is in turn driven by bull gear 14 on the shaft 15 of motor 16.

The shaft 12 for driving impeller 11 is journaled in bearings 17 and 18, which are hydrodynamic journal bearings providing full film lubrication and are fed with oil derived from an oil pump-separator 19 (which will be described later) through oil feedlines not shown. The impeller 11 is attached to the shaft 12 by an impeller retaining bolt 20 and other parts which friction key the impeller to the shaft in accordance with the general arrangement described in U.S. Pat. No. 3,601,501.

The seal arrangement according to the invention may be located in the area designated 21 in FIG. 1 and the importance of a reasonably good seal at this location separating the high pressure or scroll chamber area 10 from the interior space 22 of the sump chamber will be described in connection with FIG. 2.

In FIG. 2, the sump chamber space 22 and any space therein in communication with the space 22 is shown as being separated from the high pressure discharge space 10 by a casing wall 23 provided with seal at location 21 in the opening in the wall through which the driving shaft 12 for the impeller 11 projects.

In the particular compressor assembly design used as an example for purposes of description, the compressor discharge space 10 (FIGS. 1 and 2) in normal operation will normally have a pressure of say 65 p.s.i.g. (the undiffused static pressure at the impeller discharge), and to which the one side of the seal at location 21 is exposed, while the pressure in the sump chamber 22 will nominally be 35 p.s.i.g., with the other side of the seal exposed to this pressure. With these values, a pressure differential of 30 psi. is available to drive refrigerant gas from the discharge pressure space to the sump chamber space 22. The sump chamber space 22 is maintained at substantially the same pressure as the suction pressure at the inlet 24 at the eye of the impeller through the connecting

lines

25 and 26 which connect the oil-refrigerant separator to a fine mesh filter and oil accumulator 27 and the filter-accumulator to the suction side of the compressor, respectively. The entire refrigerant loop is hermetically closed and the lowest pressure in this loop is that at the compressor impeller inlet or suction pressure. Therefore, the leakage of refrigerant gas through the seal is required to be returned to the refrigerant system and this is accomplished by the oil pump-separator 19 which includes an electric motor 28 driving a gear pump 29 for pumping oil to the bearings in the sump chamber, and a driven oilrefrigerant separator 30 which is intended to remove large particles of oil from the refrigerant and vent the refrigerant through the line 25 to the filter and oil accumulator 27.

In normal operation of the compressor assembly the gear arrangement of the drive train generates voluminous oil mist particles which are in part entrained in refrigerant leakage into the sump chamber 22. The larger oil particles are prevented from being vented along with the refrigerant back to the refrigerant system by the centrifugal oilrefrigerant separator 30 but some of the small oil mist particles, such as those which are 2 microns or smaller, are not separated but escape through the line 25 to the filter and oil accumulator 27. This filter is provided with a filter medium which for the most part prevents the passage of those particles greater than 0.4 microns through the vent line 26 to the refrigerant system. However, these separated oil particles are stored in the accumulator portion 31 and can be drained back to the sump chamber only during compressor shutdown through the drain line 32 containing a check valve 33. In that connection it is noted that FIG. 2 is a schematic view and that the filter and accumulator 27 is located to permit gravity drain back to the sump chamber through the line 32.

The desirability of having an effective seal arises in large part from the compactness of the sump chamber. With the relatively compact sump chamber, which is desirable in connection with the overall compactness of the compressor assembly, the oil capacity in the sump chamber is limited. For example with a centrifugal compressor of the compact size having approximately a IOO-ton refrigeration producing capacity, in the particular illustrated arrangement the oil capacity is limited to approximately 2 gallons. The limited size sump chamber also imposes a disadvantage in that the inter nal surface area in the sump chamber upon which oi] mist can coalesce is relatively limited as contrasted to units with larger sump chambers. In the compact size units the total amount of oil present in the chamber is so limited that the oil which is not coalesced on surfaces is subject to being entrained by the refrigerant gas being vented back to the refrigerant system. Since the rate of oil carry-out is generally proportional to the refrigerant venting rate the result is that in a shorter time period than is desirable sufficient oil has been moved into the accumulator 27 by the venting refrigerant gas that a shutdown of the system is required to permit drainage of the oil from the accumulator back to the sump chamber. While in larger units of this general design a sump pump and venting is also provided, there is no requirement for the filter 27 because the larger oil capacity in the sump chamber permits a greater venting loss of oil and larger periods of time between shutdown of the compressor. Thus, in addition to the disadvantage of the limited time intervals between shutdowns. the compact centrifugal compressor has the further disadvantage of requiring the use of the fine mesh filter and oil accumulator device 27. From this it will be appreciated that if refrigerant leakage from the high pressure chamber to the sump chamber can be substantially reduced, the noted disadvantages can be substantially avoided. The seal arrangement of the invention is intended to accomplish this, particularly in the noted environment of a compact centrifugal refrigerant compressor. The seal arrangement is also of a reasonably low cost and does not disturb the basic design parameters of the centrifugal compressor assembly having the general impeller and shaft arrangement disclosed in U.S. Pat. No. 3,60l,50l.

In the illustrated seal arrangement of FIG. 3 the shaft 12 is supported by the hydrodynamic journal bearing 18 to which oil is fed in the direction indicated by the dash line arrow under pressure from the pump 29 (FIG. 2) to provide a full film lubrication, with the oil then passing into the oil cavity 35 which is in turn connected by means not shown to drain to the sump of the sump chamber 22. Accordingly the oil cavities are at the same pressure as the interior of the sump chamber, which in the example is approximately 35 p.s.i.g.

The seal rotor part 36 (FIG. 3) has the general shape of a circular washer with a stepped opening therein. The shoulder of the counter bored part of the stepped opening seats on the end of shaft 12. The outer circumference of the seal rotor 36 includes one portion 38 extending in a generally axial direction and provided with labyrinth teeth thereon and an outwardly, obliquely directed portion 39 also provided with labyrinth teeth thereon, these two portions being separated at an intermediate location along the total axial extent of the outer circumference of the rotor by a circumferential groove or channel 40.

The seal stator 41 (FIG. 3) has the general shape of a ring which encircles the seal rotor 36 and includes in its currently preferred form an inner ring 42 of carbon graphite which is shrunk fit in a steel outer retaining ring 43. The inner surface of the ring 42 is shaped to accommodate the profile formed by the tips of the teeth on the seal rotor. The seal stator is secured in position to the casing 23 by retaining rings 44 engaging a circumferential flange 45 of the ring 43, the flange in turn compressing an O-ring secondary seal 46 in the currently preferred construction. An anti-rotation pin 47 locking the outer section of the seal stator against rotation relative to the casing 23 may also be provided.

It will be appreciated from the description and from FIG. 3 that the dynamic seal thus provided by the seal rotor and seal stator has an axial extent which includes the one portion 38 and the other portion 39 separated by the intermediate location circumferential channel 40.

The rotor 36 (FIG. 3) has a plurality of bores 48 which extend from the channel 40 to an annular clearance space 50 defined between the shank 37 of the retaining bolt and the surrounding structure including the seal rotor 36, the hub of impeller 11, and a ringshaped spacer 49. It is noted that the impeller and shaft assembly as a whole is of the general arrangement described in US. Pat. No. 3,601,501 which results in the impeller being friction keyed to the shaft 12 and permits the clearance space 50 noted. While the size and shape of the illustrated rotor 36 dictates that the bores 48 extend obliquely, in rotors of other shapes or larger sizes they could be directly radial. In any case, the bores of this particular invention include radial components.

An axial bore 51 (FIG. 3) is provided in the retainer bolt 20 and a plurality of radially directed transverse bores 52 place the axial bore in communication with the clearance space 50 and with the rotor bores 48. Thus, a venting passage for refrigerant gas leaking into the seal from the high pressure space is provided from an intermediate location along the leakage path of the dynamic seal to the space 24 at the inlet or suction side of the impeller.

It is considered important in the particular centrifugal compressor assembly in which the seal arrangement is described as being incorporated that the labyrinth seal be manufactured with a relatively close clearance. Satisfactory results have been obtained by constructing the seal stator with an inside diameter of the carbon graphite ring being such that, as assembled to the seal rotor, a 0 to 0.002 inch diametral clearance exists between the stator and rotor. During the first few seconds of operation the sharp metal labyrinth teeth rapidly cut from the carbon what is needed, but no more. This wear-in occurs because the seal stator is clamped nonfloating) at assembly with the seal rotor as a guide. In addition the characteristics of the particular machine are such that the shaft will move radially a few mils in going from its position at assembly to its normal position at operating speed. This radial shift of the rotational centerline arises from differential thermal expansions occurring in the main bearings and from a shift of position of the rotating shaft within the journal bearings reflecting oil film thickness changes. The relatively close tolerances of the seal arrangement are considered to be important to insure that each portion of the labyrinth seal, 38 and 39 (FIG. 3) has a higher gas flow resistance than the internal vent passage leading from the intermediate location along the seal to the inlet space 24.

The way in which the seal arrangement works is as follows. In normal operation a pressure of, say, 65 p.s.i.g. exists at the impeller or high pressure side (FIG. 3) of the seal arrangement while a pressure of, say, 35 p.s.i.g. exists at the other side of the seal arrangement exposed to the oil cavity 35. The pressure in the annular groove 40 at the intermediate location will be only slightly greater than the pressure at the suction side of the impeller, which as was noted, is substantially the same as that in the oil cavity 35. For example the pressure at the intermediate location 40 at the annular groove may be in the order of 36 or 37 p.s.i.g., and with the relatively low flow resistance of the vent passage, most of the refrigerant gas leakage which reaches the intermediate location is returned to the suction side of the impeller rather than going through the relatively high resistance leakage path of the portion 39 of the seal leading to the oil cavity 35. In testing it has been determined that the ratios of reduction of leakage into (ill 6 the oil cavity with the illustrated arrangement, over the seal presently in use, ranged from about I to l0 to l to 20 depending upon the clearances between the rotor teeth and the stator.

It is currently believed preferable that the portion 39 (FIG. 3) of the seal be obliquely disposed as illustrated, rather than extending axially as the other portion 38. The reason for this will be better understood in light of the following remarks. In normal operation of the compressor the pressures are the noted nominal pressures of 65 p.s.i.g. discharge and 35 p.s.i.g. at the inlet to the impeller and in the sump chamber. However, when the compressor is shut down for a sufficient period to equalize the pressures throughout this hermetically closed system, the pressures throughout the system will be in order of about p.s.i.g. Thus upon start-up of the compressor the discharge pressure is lower than the sump pressure for a short time, such as 10 seconds. This is due to the lesser restriction of the refrigeration system as a whole relative to the restriction of the communicating lines from the sump chamber to the refrig eration system. As a result, upon start-up there can exist a reverse pressure gradient across the seal which could tend to permit oil leakage back through the seal into the refrigerant system. By providing the oblique disposition of the seal portion 39, centrifugal force in enlisted to throw-back" oil which might tend to leak back through the seal.

While the seal illustrated in FIG. 3 provides the noted significant reduction in leakage in the range of operating speeds currently used, with possible increases to higher operating speeds in the future, greater rates of leakage than occur with the current operating speeds can be expected. The projected increased leakage will result from the generally radially directed bores 48 (FIG. 3) in the seal rotor 36 producing a pumping action as if they formed a small centrifugal compressor. While this pumping action may be considered, on balance, to be advantageous to prevent reverse leakage in an installation operated normally with an appreciable number of starts and stops. this does not apply to a system intended for mostly continuous operation with infrequent shut-downs. In the illustrated arrangement the pumping action can produce a pressure of say 5 p.s.i. higher than the compressor suction pressure at the current nominal operating speeds. Because the low pressure side of the seal is at the same compressor suction pressure, there is then a 5 p.s.i. pressure difference available to drive refrigerant leakage from the annular groove 40 through the second portion 39 of the seal and into the sump chamber. While this leakage is not objectionably large at the nominal 30,000 to 32,000 rpm operating speed of the compressor, the leakage will increase with increased speeds and thus would become significantly higher with higher operating speeds. This stems from the leakage increasing with compressor speed changes from mid-range speeds, such as 10,000 to 20,000 rpm. up to and above the nominal operating speed.

The constructions which are the subject of the noted companion application of this applicant and Raimondi. are such that the adverse rising-leakage/speed characteristic at such speeds is substantially eliminated. and also yields a further reduction in leakage at the current operating speeds. Further. the constructions are such that all portions of the venting passage formed in the retaining bolt 20 are eliminated so there is no chance of weakening of the retaining bolt. Reference should be 7 had to that companion application for a disclosure of how this is accomplished.

I claim:

1. Apparatus including a casing, and a centrifugal impeller having an inlet side, the casing having an opening therein through which a shaft for driving the centrifugal impeller projects and a seal arrangement in said opening for restricting fluid leakage from an impeller discharge pressure space on one side of the seal arrangement to a second space on the opposite side of the seal arrangement maintained in normal operation of said apparatus at substantially the same pressure as the suction pressure at the inlet side of said impeller, said seal arrangement including:

a seal rotor rotating with said shaft and having an outer face;

an annular seal stator around said rotor and having a face opposing said outer face;

the opposing faces of said rotor and stator forming a dynamic seal having one end of the axial extent of the seal exposed to said discharge pressure space and the other end of the axial extent of the seal exposed to said second space;

means defining a venting passage from an intermediate location along said axial extent of said seal, said venting passage extending inwardly into said rotor and to said inlet side of said impeller, said venting passage having a lower resistance to fluid flow than the resistance to fluid flow between said intermediate location along the axial extent of said seal and either end of said seal.

2. Apparatus according to claim 1 wherein:

said dynamic seal is of labyrinth character including outwardly projecting teeth on the outer face of said rotor.

3. Apparatus according to claim I wherein:

said intermediate location is at the midpoint of said axial extent of said seal.

4. Apparatus according to claim 1 wherein:

said venting passage includes inwardly-directed discrete bores in said rotor, the inner ends of said bores being in communication with axially-extending passage means in communication with said inlet side of said impeller.

5. Apparatus according to claim 4 wherein:

said axially-extending passage means includes a bore extending along the centerline of said impeller.

6. Apparatus according to claim 1 wherein:

said seal rotor includes an annular channel extending circumferentially of said rotor at said intermediate location; and said dynamic seal includes one portion extending obliquely outwardly to said second space from said intermediate location.

7. A seal arrangement for a compressor having an impeller driven by a shaft, the impeller having an inlet side and discharging fluid received thereat to an impeller discharge space, comprising:

a seal rotor rotating with said shaft and having an outer face;

an annular seal stator surrounding said rotor and having a face opposing said outer face;

the opposing faces of said rotor and stator forming a leakage inhibiting seal having an axial extent exposed on one end to the impeller discharge space, and on the other end to a space having a pressure substantially that of the pressure at the inlet side of said impeller during normal operation;

means defining a venting passage between an intermediate location along said axial extent of said seal to the space at the inlet side of said impeller for normal flow in that direction, at least a part of said venting passage means extending radially inwardly into said seal rotor, said venting passage having a low resistance to flow relative to the resistance to flow imposed by the axial extent of said seal.

Claims

1. Apparatus including a casing, and a centrifugal impeller having an inlet side, the casing having an opening therein through which a shaft for driving the centrifugal impeller projects and a seal arrangement in said opening for restricting fluid leakage from an impeller discharge pressure space on one side of the seal arrangement to a second space on the opposite side of the seal arrangement maintained in normal operation of said apparatus at substantially the same pressure as the suction pressure at the inlet side of said impeller, said seal arrangement including: a seal rotor rotating with said shaft and having an outer face; an annular seal stator around said rotor and having a face opposing said outer face; the opposing faces of said rotor and stator forming a dynamic seal having one end of the axial extent of the seal exposed to said discharge pressure space and the other end of the axial extent of the seal exposed to said second space; means defining a venting passage from an intermediate location along said axial extent of said seal, said venting passage extending inwardly into said rotor and to said inlet side of said impeller, said venting passage having a lower resistance to fluid flow than the resistance to fluid flow between said intermediate location along the axial extent of said seal and either end of said seal.

2. Apparatus according to claim 1 wherein: said dynamic seal is of labyrinth character including outwardly projecting teeth on the outer face of said rotor.

3. APparatus according to claim 1 wherein: said intermediate location is at the midpoint of said axial extent of said seal.

4. Apparatus according to claim 1 wherein: said venting passage includes inwardly-directed discrete bores in said rotor, the inner ends of said bores being in communication with axially-extending passage means in communication with said inlet side of said impeller.

5. Apparatus according to claim 4 wherein: said axially-extending passage means includes a bore extending along the centerline of said impeller.

6. Apparatus according to claim 1 wherein: said seal rotor includes an annular channel extending circumferentially of said rotor at said intermediate location; and said dynamic seal includes one portion extending obliquely outwardly to said second space from said intermediate location.

7. A seal arrangement for a compressor having an impeller driven by a shaft, the impeller having an inlet side and discharging fluid received thereat to an impeller discharge space, comprising: a seal rotor rotating with said shaft and having an outer face; an annular seal stator surrounding said rotor and having a face opposing said outer face; the opposing faces of said rotor and stator forming a leakage inhibiting seal having an axial extent exposed on one end to the impeller discharge space, and on the other end to a space having a pressure substantially that of the pressure at the inlet side of said impeller during normal operation; means defining a venting passage between an intermediate location along said axial extent of said seal to the space at the inlet side of said impeller for normal flow in that direction, at least a part of said venting passage means extending radially inwardly into said seal rotor, said venting passage having a low resistance to flow relative to the resistance to flow imposed by the axial extent of said seal.