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Publication numberUS2975959 A
Publication typeGrant
Publication dateMar 21, 1961
Filing dateSep 30, 1958
Priority dateSep 30, 1958
Publication numberUS 2975959 A, US 2975959A, US-A-2975959, US2975959 A, US2975959A
InventorsFoster Berry W
Original AssigneeFoster Berry W
Export CitationBiBTeX, EndNote, RefMan
External Links: USPTO, USPTO Assignment, Espacenet
Back-to-back centrifugal compressor and centripetal turbine having an integral rotordisc
US 2975959 A
Abstract  available in
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Claims  available in
Description  (OCR text may contain errors)

March 21, 1961 B. w FOSTER 2,975,959

BACK-TO-BACK CENTRIFUGAL COMPRESSOR AND CENTRIPETAL TURBINE HAVING AN INTEGRAL ROTOR DISC Filed Sept. so, 1958 a Sheets-Sheet 1 INVENTOR. BERRY W I TZSTEP .AYTOKZZZ'Y March 1961 B. w. FOSTER 2,975,959

BACK-TO-BACK CENTRIFUGAL COMPRESSOR AND CENTRIPETAL TURBINE HAVING AN INTEGRAL ROTOR DISC Filed Sept. 30, 1958 3 Sheets-Sheet 2 INVENTOR. BERRY W Hostel? ATTORNEY March 1961 B w. FOSTER 2,975,959

BACK-TO-BACK C ENTRIFUGAL COMPRESSOR AND CENTRIPETAL TURBINE HAVING AN INTEGRAL ROTOR DISC Filed Sept. 30, 1958 FIG. 6 R1 5 Sheets-Sheet 3 INVENTOR. BERRY VV Fsrsn ATTORNi'Y' y i e g in.

iten in BACK-TO-BACK CENTRIFUGAL COMPRESSOR AND CENTETAL TM HAVING AN 1N- TEGRAL RGTQR DliSQ This invention relates to a novel cooling system for a unit comprising a back-to-back centrifugal compressor and centripetal turbine having an integral rotor disc.

In this invention a suitable fluid-for example, a highly heat-conductive and low-melting-point molten metal-1s hermetically sealed in a channel that passes near to the surfaces of the rotor. When the fluid is heated by hot gas on the turbine side of the rotor, the fluid expands; consequently, its density drops. This drop in density causes the hot fluid to flow radially inwardly on the side next to the turbine surface, being displaced by the heavier cooler fluid that flows radially outwardly on the side next to the compressor surface. Thus, forced thermal circulation of' the fluid conducts heat away from the turbine side of the rotor to the compressor side of the rotor.

The cooling action of the circulating fluid supplements and has a greater heat-carrying capacity than the solid metal core of the rotor disc. By having the channels for the circulating fluid strategically located, e.g., at or near the root of the turbine rotor vanes and the compressor impeller vanes, the critical parts can be kept relatively cool even though the turbine gas temperatures are high.

An additional feature of the invention calls for plating or coating the rotor surfaces.

The surface of the turbine rotor and the surfaces of its vanes are plated or coated with material which is a poor heat conductor and has reasonable strength properties at high temperatures. The effect is to reduce to a minimum the flow of heat away from the hot turbine gases. A porous material may be used for this heat-dam coating,

and it may be plated with a chromium-nickel alloy or any alloy that has good corrosion-resistant properties at temperatures. The corrosion-resistant plating may be polished to a mirror-like snrfacepto reflect away radiant heat and thus reduce the heat flow from the hot turbine gases into the turbine side of the rotor.

The surface of the compressor rotor and the'surfaces of its vanes are plated or coated with a dull black material of good heat conductivity, to increase the compressors capacity to dissipate heat to the air stream.

Other objects and advantages of the invention will appear from the following description.

n? the drawings:

Fig. 1 is a view in elevation of .a rotor embodying the principles of the invention, looking at its compressor side.

Fig. 2 is a view of the rotor, taken along the line 2-2 in Fig. l. p i

Fig. 3 is an enlarged view in section of the rotor, taken along the line 33 in Fig. l and showing the channels for the cooling fluid.

Fig. 4 is a view in section on the scale of Figs. 1 and 2, taken along the line 4-4 in Fig. 2 through the center of the rotor.

Fig. 5 is a view in section taken along the line 5-5 in Fig. 2. a

Fig. 6 is a view like Fig. 1, showing the rotor disc before the vanes are installed.

Fig. 7 is a View like Fig. 2, of the rotor disc before the vanes are installed. It may also be considered as a view looking at Fig. 6 along the direction of the arrows 77.

Fig. 8 is an exploded view, the rotor being shown turned with respect to Fig. 7 and one vane installed and three others adjacent their slots.

Fig. 9 is an enlarged side view of one Vane, taken along the line 9-9 in Fig. 8.

Fig. 10 is a view in section of one of the vanes, taken along the line 10-10 in Fig. 8.

.The drawings show a combined turbine-compressor unit 15, comprising a centrifugal compressor and a centripetal turbine placed back-to-back on an integral rotor disc 16. Rotor vanes 17 for the centrifugal compressor are anchored to one side of the rotor 16, and vanes 18' for the centripetal turbine are anchored to the other side.

The rotordisc 16 forms a ring around an axial passage 20 and is the main structure for supporting the assembled rotor parts. A feature of this invention is to keep this disc 16 at a low average temperature so that it will have a good stress life. In the radial center of rotor disc 16, at opposite ends of the axial passage 20, are small collars 21 and 22, preferably integral with the disc 16 and with inner diameters slightly smaller than the diameter of the passage 20. These collars 21, 22 may be made up of segments on the compressor and turbine blades (ref. Figs. 8 and 9), and they become integral with the disc 16 when they are welded or furnace-brazed in place.

A hollow circular shaft 23 may be pressed into the collars 21 and 2-2 and may be welded or furnace-brazed to the collars.

The annular cylindrical space between the shaft 2 3 and the axial passage it comprises a cylindrical channel 25,. The

channel 27 joins the compressor end of the channel 25.

The channels 27 may be located at or near the root of each compressor Vane 17. In the illustrative design there are twelve of these channels 27, though more or fewer may be used if desired.

Located near the surface 2% of the turbine side of the disc 16 are channels 29, which run substantially parallel to the surface 28 in a generally radial but partly axial direction. At their radially inner extremity each channel '29 joins the turbine end of the channel 25. These channels 29 may be located at or near the root of eachturbine vane 18. In the illustrative design there are twelve of these channels 29, but more or fewer may be used if desired.

Each pair of channels 27 and 29 are joined together by passages or ports 30 at their radially outer extremities.

A hole 31 may beprovided to fill the channels 25, 27, 29 V and 30 with suitable fluid 32, preferably amolten metal such as hot sodium, lithium, or mercury. A plug 33 is then inserted into the hole 31 to hermetically seal this fluid 32 in the channels 25, 27, 29, and 30.

One way in which the channels 25, 2 7, 2%, and 30may be formed will now be described as a preferred example of this structure. The compressor vanes 17 and the turbine vanes 18 may be fabricated as separate parts, as shown by Figs. 8 to 10. An enlarged anchor section-34 may be located at the root of each vane 17 and 18 and each anchor section 34 may be routed at 35 to provide part of the channels 27 and 29, the remainder of the channels being provided in the outersurface of the disc I Patented Mar. 21, i961 16 with anchor slots 36 and 37 that locate and secure ends 46 of'each pair of vanes 17 and 18 may be welded.

or furnace-brazed together, thereby providing the connecting channels 30. When the parts are welded or furnace-brazed together, care is taken to leave the channels 25, 27, 29 and 30 open, so that the cooling fluid will circulate through them.

In order to reduce the amount of heat that flows from the hot turbine gas stream between the vanes 18 to the rotor disc 16, a foamed coating 41 is applied to the turbine surface 28 of the disc 16 and to the surfaces of the vanes 3.8. This foamed coating 41 should be of a material which has good heat insulation and heat resistant properties at high temperature. For example, it may be of ceramic such as alumina or a cermet such as titanium carbide with a nickel-cobalt chromium binder. The thickness of this plating 41 may be several thousandths of an inch thick. For example, it may be 0.005" thick and may well lie in the range of 0.002" to 0.05". Thin plies of coating 41 may be attached to the disc 16 and vanes 18 by furnace brazing, by flame-spraying, or by a plating process. The coating 41 is preferably covered with a plating 42 of chromium-nickel alloy or other alloy that has a good corrosion resistance at high gas temperatures. The plating 42 may be polished to a mirror-like surface, which will reflect away radiant heat and thus reduce the heat flow from the hot turbine gases into the turbine rotor.

The compressor rotor surface 26 and its vanes 17 are preferably covered with a dull black surface coating 43 of a material with good heat conductivity, such as a dyed epoxy enamel electrostatically sprayed on an aluminum coating applied to the compressor surface by vacuum metallizing several ten-thousandths of an inch thick. This process was developed by the Perfection Finishing Corp, Wauseon, Ohio. This gives the compressor' a larger capacity to dissipate heat to the air stream.

Operation When the back-to-back centrifugal compressor and centripetal turbine unit is rotating, compressed air flows radially outwardly through the compressor impel ler vanes 17, and hot gases expand radially inwardly through the turbine vanes 18. The compressor absorbs part of the shaft energy which is produced by the gas turbine.

The hot gases flowing through the turbine rotor vanes 18 heat up thevanes 18, and the turbine side of the rotor disc 16 is heated. The cooling fluid 32 in the channels 29, which are strategically located at the roots of the vanes 18, is heated and expands, thereby decreasing its density. At the same time, the cooling fluid 32 in the channels 27, which are strategically located near the roots of the compressor impeller vanes 17, are cooled by the air passing through the vanes 17, and'this fluid 32 in the channels 27 contracts and becomes denser. flhe channels 27 and 29 are connectedat their radially outer extremities by the ports 30 and are connected at their radially inner extremities by the annular channel 25. Thus, there will be a continuous fluid circuit 27, 30, 29, 25. The high centrifugal force of the rotating disc 16 forces the fluid of highest density to flow radially out; wardly, and the fluid of lowest density to float radially nels 29 is transferred to the compressor rotor side by the flow of this hot fluid to the channels 27. Since the temperature of the compressor rotor surface 26 and the impeller vanes 17 is much lower than that of the hot liquid in the channel 27, this liquid heat is dissipated through these parts into the cold compressor air.

In addition to this liquid cooling process of the backto-back centrifugal compressor and centripetal turbine unit 15 with an integral rotor disc, there is cooling by direct conduction through the solid rotor disc 16. A well designed liquid cooling system of the type described can transfer several times the heat that is conducted through the solid disc.

The amount of heat flowing from the hot turbine gases to the rotor 16 and turbine vanes 18 is also reduced by the surface coating of these parts with the foamed coating 41, which has high thermal insulation properties. The low thermal conductivity of the foamed coating 41 restricts the amount of heat conducted to the rotor 16 and the cooling liquid in the channel 29 The surface of the vanes 18 and the turbine rotor surface 28 which face the hot turbine gases being provided with a mirror-like surface 42, most of the radiant heat is reflected away, thereby further reducing the flow of heat into the turbine vanes 18 and the turbine rotor 16. The compressor rotor surface 26 and. its impeller vanes 17 being coated with a dull black finish, the heat from the liquid channel 27 and the rotor disc 16 is dissipated more readily into the compressor air. a

By means of the unique arrangements and processes which have been described in this invention, the hot turbine gases can be at an average temperature of 2500 F., while the solid core of the rotor disc 16 can be kept at a temperature less than l500; at thistemperature a rotor of alloy steel can be designedto have a good stress life. This design thus greatly extends the range of gas turbines for the specific power and the efficiency increases directly with the temperature rise during the combustion process and the absolute temperature of the turbine gases.

To those skilled in the art to which this invention relates, many changes in construction and widely differing embodiments and applications of the invention will sug gest themselves without departing-from the spirit and scope of the invention. The disclosures and the description herein are purely illustrative and are not intended to be in any sense limiting. For example, the shape of the compressor impeller vanes is not limited to that shown, but may be swept back or otherwise shaped, if desired. After all, the exact shape of the vanes is not the essential thing in the invention. And, of course, the invention may be used in many environments other than those men-- tioned, such as for the last stage of a multistage compressor.

I claim: 7

l-. A unit comprising a compressor-turbine combination, including an annular single rotor disc, centrifugal inwardly, By this mechanism of centrifugal separation of the different densities, the hot, low-densitygliquid in the turbine cooling channels 29 flows radially inwardly, and

compressor blades on one side of said disc, and centripetal turbine blades on the other side of said disc, said unit having a cooling system comprising a cylindrical channel near the inner peripheryiofsaid disc, a series of second channels leading generally radially outwardly from the opposite sides of said cylindrical channel each said second channel being generally parallel to the surface of the side to'which it is closer and terminating at its radially -outer extremity in a port joining it to a said second channel from the opposite side, said channels being filled with heat-conductive fluid, whereby, under rotative forces, said fluid being heated on said turbine side decreases in density and flows radially inwardly as it is displaced by fluid of higher density cooled on said compressor side, and flows through said cylindrical channel to the channels on said compressor channels where its heatisdissipated.

2. The unit of claim 1 wherein said fluid is molten metal.

3. The unit of claim 1 wherein the surface of the turbine side of said disc and the turbine blades are coated with a thin coating of heat-insulating material.

4. The unit of claim 3 in which said coating is covered by a plating of a metal that is corrosion-resistant at high temperatures, polished to a mirror-like surface.

5. The unit of claim 1 in which the surface of the compressor side of said disc and the surfaces of said compressor blades are covered with a dull black coating of high heat-conductivity.

6. The unit of claim 1, in which the cylindrical channel is provided by a recessed cylindrical portion at the axis of said disc between unrecessed collar-like ends and by a cylindrical shaft extending through said disc and hermetically sealed to said collar-like portions.

7. The unit of claim 6 wherein said vanes are separate pieces from said disc and are secured thereto along slots, said series of channels being provided by recesses in said vanes adjacent said slots.

8. Means for increasing the radiant heat dissipation from a compressor rotor into a compressor air stream; comprising bonding to said compressor rotor impeller vanes and said compressor rotor surface with a material of good heat conductivity to provide a coating with a thin dull black finish; said dull black finish increasing the heat flow from the rotor to the compressor air stream and thus lowering the rotor temperature on the compressor side, said rotor disc thereby operating at a lower average temperature due to the increase of heat dissipation into the air stream.

9. Means for cooling a back-to-back centripetal turbine and centrifugal compressor with an integral rotor disc having turbine rotor vanes on one side and compressor impeller vanes on the other side, by the thermal and centrifugal forced circulation of a cooling liquid, comprising means defining a liquid channel at the root of each turbine rotor vane, means defining a liquid channel at the root of each compressor impeller vane, means connecting said liquid channels at the inner radius of said disc, and means connecting opposite pairs of said channels at the 'outer radius of said disc, said channels thereby being part of a loop circuit around a solid disc section, a liquid of high heat conductivity hermetically sealed in said channel circuit so that when the rotor is rotating under operating conditions hot turbine gases heat said liquid at the roots of said turbine rotor, vanes causing said heated liquid to expand and reduce its density, the cold compressor air meanwhile cooling said liquid at the roots of said compressor rotor vanes; said cooled liquid thereby contracting and increasing its density; the centrifugal field due to rotation thereby forcing the cooler higher density liquid to flow toward the outer periphery of said disc, and the hotter lower density liquid floats to the inner periphery of said disc.

10. An integral rotor disc comprising a unitary core with a central core hole and anchor slots on opposite sides, separate compressor rotor vanes each with an enlarged anchor portion at its root, separate turbine rotor vanes each with an enlarged anchor portion at its root, all said enlarged anchor portions having channels therein, said anchors being held in place in said anchor slots; a cylindrical shaft in said core hole engaging said disc at its outer ends only to provide a central annular channel therebetween, and a cooling liquid of high heat conductivity at the operating temperatures filling said channels and hermetically sealed therein.

References Cited in the file of this patent UNITED STATES PATENTS 2,447,292 Van Acker Aug. 17, 1948 2,641,440 Williams June 9, 1953 2,709,893 Birmann June 7, 1955 FOREIGN PATENTS 677,052 Great Britain Aug. 6, 1952

Patent Citations
Cited PatentFiling datePublication dateApplicantTitle
US2447292 *Oct 12, 1943Aug 17, 1948Joseph E Van AckerGas-actuated turbine-driven compressor
US2641440 *Nov 18, 1947Jun 9, 1953Chrysler CorpTurbine blade with cooling means and carrier therefor
US2709893 *Aug 6, 1949Jun 7, 1955Laval Steam Turbine CoGas turbine power plant with heat exchanger and cooling means
GB677052A * Title not available
Referenced by
Citing PatentFiling datePublication dateApplicantTitle
US4155151 *Aug 31, 1977May 22, 1979Stiegelmeier Owen EHeavy duty impeller and method of fabricating the same
US4322200 *May 7, 1979Mar 30, 1982Stiegelmeier Owen EHeavy duty impeller
US5878808 *Oct 30, 1996Mar 9, 1999Mcdonnell DouglasRotating heat exchanger
US7438529Apr 6, 2006Oct 21, 2008Daimler AgCompressor and turbine wheel for a secondary air feed device
DE102006048784A1 *Oct 12, 2006Apr 17, 2008Man Diesel SeVerdichter für einen Turbolader sowie Verfahren zu dessen Kühlung
WO2005040565A1 *Sep 8, 2004May 6, 2005Daimler Chrysler AgCompression and turbine wheel for a secondary air conveying device for an internal combustion engine
WO2005040573A1 *Sep 8, 2004May 6, 2005Daimler Chrysler AgCompressor and turbine rotor for a secondary air supply device on an internal combustion engine
Classifications
U.S. Classification416/96.00R, 415/114, 416/184, 165/86, 165/104.25, 416/175
International ClassificationF02C3/05, F01D5/04, F01D5/02, F02C3/00, F01D5/28
Cooperative ClassificationF01D5/04, F01D5/046, F02C3/05, F01D5/28
European ClassificationF01D5/04C3, F01D5/28, F02C3/05, F01D5/04