US3765480A

US3765480A - Device for cooling rotors

Info

Publication number: US3765480A
Application number: US00135606A
Authority: US
Inventors: P Fries
Original assignee: Siemens AG
Current assignee: Siemens AG
Priority date: 1970-04-24
Filing date: 1971-04-20
Publication date: 1973-10-16
Anticipated expiration: 1990-10-16
Also published as: JPS5214452B1; AT306167B; DE2019956A1; SE361794B; CH528170A; CA918730A

Abstract

A cooling device for cooling a rotor comprises a condensation chamber on a front surface of the rotor and a plurality of cavities containing an evaporable working medium provided in the area of the heating zones at the periphery of the rotor and ending directly in the condensation chamber.

Description

United States Fries tent DEVICE FOR COOLING ROTORS Inventor: Paul Fries, Erlangen, Germany Assignee: Siemens Aktiengesellschaft,

Berlin, Germany Filed: Apr. 20, 1971 Appl. No.: 135,606

[30] Foreign Application Priority Data Apr. 24, 1970 Germany P 20 19 956.8

References Cited UNITED STATES PATENTS 1,739,137 12/1929 Gay ..3l0/54X Get. 16, 1973 2,794,135 5/1957 Swendsen 165/86 X 2,799,259 7/1957 Famy 61 al 165/86 X 2,782,000 2/1957 Ledinegg 416/96 2,330,121 9/1943 Heintz 165/104 X Primary ExaminerAlbert W. Davis, Jr. Attorney-Curt M. Avery, Arthur E. Wilfond, Herbert L. Lerner and Daniel J. Tick ABSTRACT A cooling device for cooling a rotor comprises a condensation chamber on a front surface of the rotor and a plurality of cavities containing an evaporable working medium provided in the area of the heating zones at the periphery of the rotor and ending directly in the condensation chamber.

9 Claims, 10 Drawing Figures 1b 7+1 w 11 s ,5

//1//// i ///1///t///y// /fl 5a i" 7 9 E! R L 4 1 I T ///1/////////////////// PAIENTED OCT 16 I975 SHEET 10F 3 DEVICE FOR COOLING ROTORS DESCRIPTION OF THE INVENTION The invention relates to a device for cooling rotors. More particularly, the invention relates to a device for cooling rotors with a heat tube containing an evaporable working medium and wherein the centrifugal acceleration may be utilized for recycling the condensate into the evaporation zone. Cavities containing the working. medium are provided within the area of the heating zones, at the periphery of the rotors.

A publication of the American Society of Mechanical Engineers, New York, Aug, 1969 (ASME Publication 69-HT-l9) of H. Gray, entitled The Rotating Heat Pipe A Wickless, Hollow Shaft for Transferring High Heat Fluxes, discloses an arrangement or device for cooling rotors, where the axis of rotation of the rotors is developed as a heat tube containing a volatile working medium. During rotation, the wall of the tubular heater is lined with the working medium due to centrifugal force. The working medium evaporates in an operational state, is condensed in a condensing chamber which is situated outside the rotors at the shaft ends, and flows back into the evaporation chamber, also due to contrifugal force. The heat due to energy losses is removed, together with the evaporated working medium, from the rotor to the condensation chamber from which it passes to the outside.

The aforedescribed type of arrangement is well suited, for example, for cooling electrical mechinery. Difficulties are encountered in large rotors particularly in rotors having larger diameters, especially when the heat due to energy losses occurs in heating zones located at the periphery of the rotor where the heat accumulates to high values. In this case, the dissipation of heat via the rotation shaft entails a high heat resistance, to which the flow of heat in the rotor is exposed. Moreover, the heat-transferring wall of the heat tube may become too small for heat transformation alone. Therefore, arotor equipped with the known cooling device does not always insure a dissipation of heat losses which is adequate, for example, for the operational reliability of an electrical machine.

A device for cooling turbine blades is known, for example, from the known publication, where the blade rotor is designed as a heat tube and cavities are provided in the turbine blades. The cavities are connected to the tubular heater via a tube having a small cross section. During operation, the centrifugal force causes the wall of the heat tube to be lined with the working medium and the working medium is forced into the cavities of the turbine blades. The working medium is primarily evaporated in the cavities of the turbine blades. The heat transfer from the turbine blades to theheat tube occurs via a flow of gas bubbles of the working medium and through a natural convection of the heated working medium in the connecting tubes. The transfer of heat via gas bubbles and the heated working medium must be carried out against the flow of the cold working medium, in these connecting tubes. The travel path of the heat is therefore also subjected to a great heat resistance. Also, when the heat losses are greater, the heat transfer area of the heat tube becomes too small. Hence, the indicated problems cannot be solved by the known arrangements.

It is the object of the invention to provide a device of the aforedisclosed type which provides adequate cooling of the rotor even when there are high heat losses.

To accomplish the aforementioned object, and in accordance with the invention, the cavities end directly in a condensation chamber located at a front surface of the rotor.

The cavities may be provided in the form of channels. It is preferable to interconnect the cavities in the rotor by means of channels. A rotor or a number of such rotors comprising one or more spaced rotor parts affixed to a shaft which functions as an axis of rotation, may be provided with cavities whereby the cavities of adjacent rotor components or of the rotors may be connected through tubular connections, or each cavity or several interlinked cavities may be directly connected to a condensation chamber located at the front surface of the rotor or rotor part or component.

Each front surface of a rotor may be provided with a condensation chamber. Preferably, each condensation chamber has a heat exchanger whose heat exchange area is larger than the area on the front surface of the rotor covered by the condensation chamber. Each condensation chamber may at least cover almost one front surface of the rotor.

The heat exchanger is preferably provided as a folded surface. The folded surface may be designed as a bellows having essentially circular concentrically positioned folds. In a bellows-type folded surface, the opening angle of the bellows folds may be made a maximum through the heat exchanger surface, at a predetermined rotor speed, with respect to the entire flow of heat.

A number of condensation chambers may be provided, and may be interconnected in the form of a cascade in order to enlarge the heat exchange area.

In the device of the invention, the evaporation of the working medium is effected directly at the place of origin, in the heating zones of the rotor. Hence, the heat due to energy losses is not deflected to the shaft and heat resistances in the flow path of the heat are therefore eliminated in the rotor. The working medium which evaporates in the cavities of the rotor flows out in the form of steam, directly into the condensation chambers at the front surfaces of the rotor. Here, too, no resistance is encountered in the flow path of the steam. The condensate, too, flows directly from the condensation chambers back into the evaporation chambers of the rotor, that is, along the outside of the connecting paths, because of the centrifugal force.

It should be pointed out that cavities cross-linked with channels will provide the rotor with separate flow paths for the steam and for the condensate. Moreover, such cross-linking may also insure the cooling of several rotors or rotor parts or components, without any additional expenditure. This type of cross-linking also helps to adjust or adapt the arrangement of the cavities in the rotor to special temperature profiles. For example, one of the cavities for the working medium in a poleshoe of an electrical machine may be guided to a location where particularly high losses occur.

The device of the invention may also be conceived, according to this description, as a system comprising various heat tubes connected within the rotor in parallel with respect to the heat flow and having common condensation chambers.

It must be emphasized, however, that the size of the condensation surfaces and the heat transfer surfaces of the heat exchanger may be adjusted in the simplest ing heat due to energy losses. The size of these surfaces is no larger determined, as in the known devices, by the diameter of the shaft, designed as a tubular heater. Hence, adequate cooling may be obtained in any case with the device of the invention, whereby cascade type connected condensation chambers may be used for extremely high heating. Shaft portions located outside the rotor may be hollow and may be included in a condensation chamber. If necessary, the shaft of the rotor, which is intended to function as the axis of rotation, may also be designed as a separate heat tube.

In order that the invention may be readily carried into effect, it will now be described with reference to the accompanying drawings, wherein:

FIG. 1 is a view, partly in section, of an embodiment of the cooling device of the invention;

FIG. 2 is a sectional view taken along the lines II-II of FIG. 1;

FIG. 3 is a schematic diagram explaining the theory of operation of the embodiment of the cooling device of FIGS. 1 and 2; 7

FIG. 4 is a view, partly in section, of another'embodiment of the cooling device of the invention,

FIG. 5 is a cross-sectional view through part of the cooling device of FIG. 4;

FIG. 6 is a view, partly in section, of another embodiment of the cooling device of the invention;

FIG. 7 is a sectional view of still another embodiment of the cooling device of the invention;

FIG. 8 is a sectional view of another embodiment of the cooling device of the invention;

FIG. 9 is a sectional view of another embodiment of the cooling device of the invention; and

, FIG. 10 is a schematic diagram, partly in section, of a flywheel Diesel generator cooled by the cooling device of the invention.

In the FIGS., the same components are identified by the same reference numerals.

FIG. I shows a rotor 1, in section. The rotor 1 comprises metal and may constitute the rotor of an electrical machine. The bearings of the rotor 1, as well as other parts of the device such as, for example, the stator of the electrical machine, are not shown in the FIGS., in order to preserve the clarity of illustration. The rotor 1 may be rotated about or around a shaft 2, which is positioned on a main inertia axis of the rotor I. The shaft 2 has a diameter 5. I It is assumed that in the rotor 1, the heat due to energy losses occurs mainly at the periphery. Therefore, a heating zone la is assumed for the rotor 1, and is separated from the remaining part 1b of the rotor, as indicated by broken line 3. The broken line 3 may be defined by points of the same temperature.

In a device of the prior art, the heat losses should be deflected, as previously stated, from the rotor portion lb to the shaft 2. The disadvantages resulting from this are hereinbefore discussed in detail. In the rotor l of FIG. 1, cavities 4 are provided in the heating zone 1a. The cavities 4 end directly in

condensation chambers

5 and 5 provided on both front surfaces of the rotor I. In the embodiment of FIG. 1, the cavities 4 are channel-type bores having longitudinal axes essentially parallel to the axis of rotation of the rotor 1.

Each condensation chamber 5 and 5' covers a corresponding front surface of the rotor l and is provided with a

heat exchanger

6 and 6, respectively, comprising a material having good thermal conductivity such as, for example, aluminum, whose heat exchange surface is larger than the corresponding front surface of the rotor 1. Any heat exchanger which is adapted to the problem may be used as the heat exchanger for the condensation chambers 5 and 5'. It was found to be particularly favorable to use a folded surface or bellows-type structure as each of the

heat exchangers

6 and 6. The shape of the cross-section and the surface of the folds or creases may be selected as desired for this type of folded surface. The folded surface may be meandershaped, wave-shaped or bellows-shaped.

FIG. 1 shows the heat exchangers 6 and 6' in form of bellows. The folds of the

bellows

6a, 6b and 6c and 6a, 6b and 6c are circular and positioned concentrically with the axis of rotation of the rotor 1. The diameter of the outer limit of the larger bellows fold is R. The openthrough the heat exchangers 6 and 6', which are hereinafter discussed in greater detail.

A separate cooling must be provided for the outside of the folding surface of each of the

heat exchangers

6 and 6. The outside cooling is schematically indicated for the heat exchanger 6' by arrows 7 and 7'. Thus, for example, each of the

heat exchangers

6 and 6 may be wetted with an air current or with a liquid coolant such as, for example, water.

The condensation surfaces of the heat exchangers 6 and 6', as shown in FIG. 1, must be mechanically protected against radial expansion due to centrifugal force. Hence,- an expansion protection ring 8 is provided on the heat exchanger 6 of FIG. 1. The folds of the condensation areas of the heat exchanger 6 are mutualy supported by expanding rings 9, which are provided with inlet openings 9a for the coolant. The expanding rings 9 may also assist in converting a laminar coolant flow into a turbulent current. This improves the heat exchanger.

The

condensation chambers

5 and 5 and the cavities 4 of the rotor l, which function as evaporation chambers, contain a working medium having a high evaporation heat and whose steam pressure is below the pressure limits, resistance or stability of the rotor when said rotor is at a permissible heating temperature. The working medium must also have chemical stability in order to prevent the formation of a gas pocket or cushion, for example, in the

condensation chambers

5 and 5.

In a state of rotation, that is, during the rotation of the. rotor l, centrifugal force causes the walls of the cavities or evaporation chambers 4 which are turned toward the periphery of the rotor 1 to be coated or lined with the working medium. The working medium evaporates and the steam flows from the evaporation chambers or cavities 4 into the condensation chambers 5 and 5'. The flow of steam is characterized in FIG. 1 by the broken arrows l0 and 10', respectively. In the condensation chambers 5 and 5', the steam condenses at the heat transfer surfaces of the

heat exchangers

6 and 6, and the heat due to energy losses produced in the rotor 1, is removed to the outside, as a result. The condensate flows back into the evaporation chambers or cavities 4.

The evaporation chambers 4 and the condensation chambers 5 and 5' thus define a hermetically sealed system, wherein the working medium circulates or cycles as a heat transfer medium, under the influence of the centrifugal force. The return flow of the condensate occurs also at the peripheries of the

condensation chambers

5 and 5 and the rotor l. The condensate flowing back into the cavities 4 is depicted in FIG. 1 by arrows l1 and 11'. A separation of the steam flow and 10' and the returning or recycling condensate l1 and 11 occurs in the aforedescribed device due to cent'rifuga] force. Hence, resistances in the flow path of the heat used to deflect the heat losses, almost never occur in the described device, since a flow of heat is prevented across the rotor l, and the steam flow 10 and 10' and the condensate flow 11 and 11' are separated. Furthermore, the

heat exchangers

6 and 6 have very large heat transfer areas available for the transforma tion or transfer of heat. Thus, the cooling device of the embodiment of FIG. 1 offers the advantages hereinbefore explained in detail.

FIG. 2 shows a section through the rotor of FIG. 1. FIG. 2 shows that, in the illustrated example, the evaporation chambers or cavities 4 are distributed in rotational symmetry, relative to the axis of,rotation of the flow Q, via the heat exchange areas or surfaces of the heat exchangers 6 and 6'. To provide optimization, we use a theory by W. Nusselt, which appeared under the title Die Oberflachenkondensation des Was serdampfes, or Surface Condensation of Steam, in the periodical of the Society of German Engineers, Vol.

60,1916, pages 541 to 546 and 569 to 575, which theory should be modified according to the problem on hand. Nusselt calculates the condensation coefficient, the thickness of the developing condensation film'and the heat flow during the condensation of flowing steam at a vertically positioned wall of constant width, under the influence of centrifugal force. The calculation is derived from the Navier-Stokes movement equations 'for a viscous liquid, which are modified under limiting con ditions and whose integration leads to the indicated quantities.

In order to apply the Nusselt water skin theory to the condensation problem on hand, that is, to a bellowstype folded surface with circular, concentric bellows folds of conical cross-section, FIG. 3 shows that the

folding surface

60, 6b and 6c of the heat exchanger 6 is reduced to a frustum housing area 12. This approximation makes sense, since, during rotation, the steam and the condensate are pressed against the

surfaces

13a, 13b and 130 of the bellow folds 6a, 6b and 6c subjected to centrifugal force. Hence, a condensate skin or coating forms at these surfaces. A statistically distributed drop condensation may develop at the surfaces 14a and 14b of the bellows folds of FIG. 3 during the condensation under the influence of the centrifugal force. The drops are flown off centrifugally. The heat transfer through the surfaces 14a and 14b is hampered only by a negligible condensate skin. This heat transfer remains uneffected by the optimization, and only the contribution of the bellows surfaces 13a, 13b and 13c which are exposed to the centrifugal force need be considered. The conical housing surface 12 of FIG. 3 is a combination of the bellows surfaces.

In the conical housing surface or area 12 which rotates around the axis of rotation at a speed f in rpm, the centrifugal acceleration b increases with the radius r, and thereby also its tangential acceleration component b,-, along a housing line of said conical housing surface. Hence, as the distance from the axis of rotation 15 increases, the condensate is subjected to increasing forces in wall direction, which try to drive the condensate to the outside. Also, an increasing radius r causes an increase in the circumference of the conical housing surface 12. The constant ground acceleration g which enters into Nusselts theory must be replaced by the variable conical housing circumference, via the variable tangential acceleration component h and via the constant wall width of the perpendicularly positioned wall, according to Nusselts theory.

Taking the coordinate system (x,y, 'y) indicated in FIG. 3, as a basis, the layer thickness of the condensate skin and the heat transfer factor a (x), during centrifugal condensation in a conical chamber, may be calculated in analogy with Nusselts theory. Both quantities are dependent upon x and y. The heat transfer factor a (x) of the condensate, the heat or thermal conductance coefficient of thematerial of the frustum wall, as well as the heat transfer factor a, on the outer surface of the cone, may help to define the median heat transfer factor along the conical condenser. From such, we may determine the heat flow Q through the areas which are coated by the condensate skin.

Assuming that the wall thickness of the conical housing 12 is very small relative to the thermal conductance coefficient of the material of the conical wall, we obtain for Q:

'wherein T, is the saturation temperature of the steam, T is the coolant temperature at the outer wall of the frustum and T,,, is the temperature of the conical wall, w [cal/gr] is the evaporation heat, p gr/cm is the density of the condensate, and 1; gr/cm is the dynamic viscosity of the condensate.

Since only the

surfaces

13a, 13b and figure in the calculation, integration should be effected only through the limits r, which are derived from FIG. 3. This equation results in the optimum opening angle, with Compared to angles which are in the vicinity of the optimum value, an increase in the heat flow up to several percent may be obtained at a given speed and given material constants, with the optimized angle.

FIG. 4 shows an embodiment, where the evaporation chambers or cavities 4 in the rotor l are interconnected through cross-linking channels 4a, which end within the rotor. The interlinking of the cavities 4 in the rotor 1 defines an additional branch system for the steam flow within the rotor l. The position of the evaporation chambers 4 and of the cross-linking channels 4a depends upon the constructive factors such as, for example, an anisotropic temperature distribution in the rotor 1. In this embodiment, too, the evaporation chambers 4 andthe channels 4a may be in rotational symmetry. The cross-linking increases the cross-section of the path of the steam flow and, as hereinbefore described, the cross-linking provides an almost complete separation of the steam flow paths 10 and 10' and the

condensate paths

11 and 11 which flow back into the evaporation chambers 4. FIG. 4 indicates that the cross-linking may continue into the area of the shaft 2.

FIG. 5 shows in section, a cross-section through the rotor l of FIG. 4. Evaporation channels or cavities 4 are located in the area of the heating zone la. The longitudinal axes of the cavities 4 need not be parallel to the axis of rotation of the rotor I but, as shown in FIG. 5, may also proceed in radial direction and the cavities may have cross-connections. The arrangement of the cavities 4 must be selected so that in a given usage, an optimum heat transfer from the heating zone la will be insured. The optimum heat transfer may be additionally controlled by the volume of working medium present in the system.

As shown in FIG. 5, the amount of working medium should be so selected in order to achieve an optimum heat transfer, that during operation, that is, during the rotation of the rotor l, a liquid level 16 is established in the cavities 4 and the cross-linked channels 4a, which is approximately within the range of the limit or boundary 3 of the heating zone la of the rotor 1. The evaporation of the working medium is then effected at this liquid level and the steam flows off through the cavities which are not filled with liquid or through the channel portions into the condensation chambers 5 and 5'. Thus, a possibly obstructing counterflow of steam and condensate in the same connecting line may be largely prevented. Instead, the steam flow paths l0 and and the condensate flow paths Ill and 11 have their own systems. During the standstill of a rotor, as illustrated in F I08. 4 and 5, the condensate collects at the deepest point of the channel or cavity system 4 and is then flown off again, during rotation, as shown in FIG. 5, to the periphery of the rotor l.

FIG. 6 illustrates an embodiment in which

several rotors

17, 18, 19, 20 and 21 are affixed to the shaft 2. The rotor 17 has two

separate rotor parts

17a and 17b. The device of FIG. 6 rotates around the axis of rotation 15 of the shaft 2. Each of the rotors 17 to 21 and each of the

rotor parts

17a and 17b is provided with cavities 4, which may be cross-linked with channels, as shown in FIGS. 1 to 5. The cavities or channel systems 4 are not separately shown in the rotors of FIG. 6. Through tubular lines 4b, the channel systems or cavities 4 of each rotor are connected to a channel system of an adjacent rotor or directly to the condensation chambers 5 or 5', provided in the embodiment of FIG. 6 at the front surface of outside rotors of the system, via a plurality of tubes 4b. The condensation chambers 5 and 5' may also be provided on the front'surface of a rotor part or component, or a rotor within the system, for example,

between the

rotor parts

17a and 17b, or at front surfaces of the

rotors

21 or 22. This system is used, for example, for cooling gas turbine blades or vanes, so that the system of FIG. 6 does not have the disadvantages of the aforedescribed known cooling system, and is not limited with respect to the quantity of heat due to energy losses, which can be removed by the cooling system.

In systems with an extremely high development of heat due to energy losses, it may become necessary to further enlarge the heat transfer surface of the

condensation chambers

5 and 5. A possibility for enlarging the heat dissipation surfaces is shown in FIG. 7. The total arrangement is disclosed with reference to FIG. 1. However, in this embodiment, the shaft portions 2a and 2b are designed as hollow or tubular

shafts having chambers

22a and 22b, respectively, formed therein. The

chambers

22a and 22b in the truncated shafts 2a and 2b are connected via

openings

23 and 23 to the

condensation chambers

5 and 5. Heat exchangers 24 and 24' are mounted on the truncated shafts 2a and 2b, respectively. Hence, in the device of FIG. 7, the

chambers

22a and 22b of the shaft 2 are incorporated into the

condensation chambers

5 and 5, thereby enlarging the heat transfer surface for the removed heat due to energy losses.

Another peculiarity of the embodiment of FIG. 7 is that the truncated shafts 2a and 2b have an approximately constant surface temperature. This may sometimes result in an impermissibly high heating of the bearings, of which a bearing 25 is schematically shown in FIG. 7. If the impermissibly high heating of the bearing cannot be obviated by a suitable bearing construction, the bearing 25 must be thermally insulated, directly at its seat, against the shaft 2. To accomplish this, a sleeve 26 ofa material having poor thermal conductivity is inserted on the shaft 2'. The sleeve 26 should be arranged so that it projects sufficiently beyond both sides of the bearing 25. The sleeve 26 may also be located in an annular depression of the shaft 2'.

Another possibility for expanding the heat transfer surface for the removal of dissipated heat or heat losses is shown in the embodiment of FIG. 8. In FIG. 8, a condensation chamber 5a is positioned on the front surface of the rotor which is shown in section.

Additional condensation chambers

5b and 5c are fixed to the shaft 2'.

The

condensation chambers

5a, 5b and 5c are interconnected in the form of cascades via a pipeline system.

The heat exchangers 6A, 6B and 65C of the

condensation chambers

5a, 5b and 50, respectively are separately wetted by a coolant flowing in the direction of arrows 7. The steam may flow through connecting lines 27 and through the chamber 23 of the shaft 2, to the

condensation chambers

5a, 5b and 5c.

The connecting lines 27 for the steam preferably branch off from surfaces of the

heat exchangers

6A, 6B and 6C, which are not affected by any component of the centrifugal force, as hereinbefore discussed in connection with FIG. 3. The condensate flows back into the evaporation chambers 4 in connecting lines 27b located at the periphery of the rotating system. The

individual condensation chambers

5 and 5 may be constructed differently, so that the cooling would be particularly enhanced. Thus, the folded surface of the heat exchanger 6C of FIG. 8 has a wavy cross-section, deviating from the folded surfaces of the heat exchangers 6A and 6B. Tubular systems may also be used as the condensation chambers, and it is particularly advantageous to cross-link the tubes of a system.

If heat losses occur not only at the periphery of the rotor l, but also within the range of the shaft in the rotor, it may be necessary to cool such area of the rotor separately. To accomplish this, as illustrated in FIG. 9, the shaft 2" may be designed as a separate heat tube. The shaft 2" must then be provided with a bore 28 which extends through the rotor and which contains a heat exchange medium. It may be necessary in the embodiment of FIG. 9 to provide the bearings for the rotor with heat insulation, as described with reference to FIG. 7.

It has been hereinbefore mentioned in this connection that several separate systems, disclosed in FIGS. 1 to 5, may be situated in one rotor. At the same time, each of the systems must be provided with its own condensation chamber on a front surface of the rotor. By utilizing such separate systems, we may obtain, for example, a layer-wise, superimposed arrangement of cooling zones in the rotor.

FIG. l shows, as an example, the cooling of the rotor of a flywheel Diesel generator by the cooling device of the invention. A generatorshaft 29 is rotatably mounted in

bearings

30 and 30, affixed to foundations 31 and 3l'respectively. A rotor 32 is mounted on the generator shaft 29. The shaft 29 drives an exciter machine 33. Pole shoes 35 are mounted with windings 36 on a rotor ring 34. The generator hasa stator 37 which is enclosed by a generator housing 38. To obtain the torque required for operation, the rotor 32 is provided with a flywheel 39.

The flywheel 39 is a so-called lateral flywheel, which is pulled up in radial directions on the driving side, by the rotor ring 34 in order to provide the shortest possible structural length. The lateral flywheel 39 is very well suited for the installation of a cooling device for the pole shoes 35, in accordance with the invention.

As shown in FIG. 10, the lateral flywheel 29 is sealed by a heat exchanger 6, which in the example of FIG. 10 has a folded surface of wave-shape. Another heat exchanger such as, for example, a star-shaped, hollow rib arrangement, may be provided instead of the folded surface. Due to the size of the flywheel 39, a condensation chamber is provided in any event. The heat transfer surfaces of the heat exchanger 6 are sufficiently large for the removal of high heat due to energy losses.

The pole shoes 35 are provided with a cross-linked system of channels, as disclosed with reference to FIGS. 4 and 5. The channel system of the pole shoes 35 is not separately shown in FIG. 10. The channel system of the pole shoes 35 is directly connected, via

tubes

40a and 40b to the condensation chamber 5. The condensation chamber 5 and the channel system contain an easily evaporable working medium. The channel system, as an evaporation chamber, and the condensation chamber 5 constitute an hermetically sealed system, wherein the working medium circulates as a heat carrier medium under the influence of centrifugal force, in the aforedescribed manner, and wherein the heat due to energy losses are transferred from the pole shoes 35 to the heat exchanger 6.

The steam flow paths l0 and condensate return flow path 11 end separately in the

tubes

40a and 40b. The outside surface of the folded surface heat exchanger 6 is wetted by a coolant 7. As a turbulence aid for the coolant flow, expanding rings may be provided for the folds of the heat exchanger, as disclosed in FIG. 1. The

expanding rings have openings for the flow of the coolant. The expanding rings protect the condensation surfaces against radial bulging due to centrifugal force. The expanding rings are not shown in FIG. 10.

The example of FIG. 10, in particular, shows what simple means may be utilized, in accordance with the invention, to provide cooling for a rotor l or 32, along its periphery. An economical, low output is required for removing even extremely high dissipation losses and the operational reliability of electrical machinery may be considerably improved thereby.

While the invention has been described by means of specific examples and in specific embodiments, I do not wish to be limited thereto, for obvious modifications will occur to those skilled in the art without departing from the spirit and scope of the invention.

I claim:

1. A cooling device for cooling a rotor rotatable about its axis and having two front faces comprising means defining a'plurality of cavities disposed at the periphery of said rotor, each of said cavities having at least a portion thereof extending substantially parallel to the axis of said rotor, a condensation chamber on a front face of said rotor, each of said cavities communicating with said chamber, said chamber and said cavities defining a hermetically sealed enclosure in which an evaporable working medium is disposed, said condensation chamber being defined at least in part by folded heat exchange surface having an area which is greater than the area of the front face of the rotor on which said condensation chamber is disposed, said heat exchange surface being defined by a plurality of substantially circular and concentrically positioned folds, the outside of said heat exchange surface being exposed to a heat exchanger medium to cool the working medium in said hermetically sealed enclosure, the inside of said heat exchange surface providing a condensation zone for said working medium, whereby centrifugal force is utilized for providing return flow of the condensed working medium from the condensing chamber to said plurality of cavities at the periphery of said rotor where said working medium is evaporated.

2. A cooling device as claimed in claim 1, wherein the cavities are formed as channels.

3. 'A cooling device as claimed in claim 1, wherein the cavities are distributed in substantially rotational symmetry to the axis of the rotor.

4. A cooling device as claimed in claim 1, further comprising a plurality of channels connected in a network with the cavities.

5. A cooling device as claimed in claim 1, further comprising a plurality of channels cross-linked in a network with the cavities in a manner whereby a level for a liquid is adjusted during operation to the volume of working medium in the network, the level being substantially within the limit of the heating zones of the rotor.

6. A cooling device as claimed in claim 1, wherein the rotor comprises an axially extending shaft and a plurality of spatially separated rotor components and rotors affixed to the shaft for rotation about the axis of the rotor, the condensation chamber is on a front surface one of a rotor component and rotor, the cavities are provided in each rotor component and rotor, and further comprising a plurality of tubular connections connecting the cavities of next-adjacent rotor components and rotors whereby each cavity and plurality of interlinked cavities end directly in the condensation chamber.

7. A cooling device according to claim 1 including a second condensation chamber on the other front face of said rotor, each of said cavities communicating with said second condensation chamber, said second chamber being included in said hermetically sealed enclosure in which said evaporable working medium is disposed, said second condensation chamber being defined at least in part by a second folded heat exchange surface having an area which is greater than the area of the front face of the rotor on which said second condensation chamber is disposed, said second heat exchange surface being defined by a plurality of substantially circular and concentrically positioned folds, the outside of said second heat exchange surface being exposed to a heat exchange medium to cool the working medium in said hermetically sealed enclosure, the inside of said second heat exchange surface providing a condensation zone for said working medium, whereby contrifugal force is utilized for providing return flow of the condensed working medium from said second condensing chamber to said plurality of cavities at the periphery of said rotor where said working medium is evaporated.

8. A cooling device for cooling a rotor having a heat tube containing an evaporable working medium and utilizing centrifugal acceleration for the return flow of the condensate into the evaporation zone, the rotor having an axis and being rotatable about its axis and having a pair of spaced front surfaces substantially perpendicular to the axis, said cooling device comprising a condensation chamber on each front surface of the rotor, a plurality of cavities containing the working medium provided in the area of the heating zones at the periphery of the rotor and ending directly in the condensation chamber, a pair of heat exchangers each provided for and cooperating with a corresponding one of the condensation chambers, each of the heat exchangers having a heat exchange surface greater than the area of the corresponding one of the front surfaces covered by the corresponding one of the condensation chambers, each of the heat exchangers comprising a folded surface in the form of a bellows, each folded surface comprising a plurality of substantially circular concentrically positioned bellows folds.

9. A cooling device as claimed in claim 8, wherein the bellows folds are coaxial and concentric with the axis of the rotor and the opening angle 2 y of the bellows folds complies with the equation tance coefficient of the material of the frustum wall.

Claims

1. A cooling device for cooling a rotor rotatable about its axis and having two front faces comprising means defining a plurality of cavities disposed at the periphery of said rotor, each of said cavities having at least a portion thereof extending substantially parallel to the axis of said rotor, a condensation chamber on a front face of said rotor, each of said cavities communicating with said chamber, said chamber and said cavities defining a hermetically sealed enclosure in which an evaporable working medium is disposed, said condensation chamber being defined at least in part by folded heat exchange surface having an area which is greater than the area of the front face of the rotor on which said condensation chamber is disposed, said heat exchange surface being defined by a plurality of substantially circular and concentrically positioned folds, the outside of said heat exchange surface being exposed to a heat exchanger medium to cool the working medium in said hermetically sealed enclosure, the inside of said heat exchange surface providing a condensation zone for said working medium, whereby centrifugal force is utilized for providing return flow of the condensed working medium from the condensing chamber to said plurality of cavities at the periphery of said rotor where said working medium is evaporated.

3. A cooling device as claimed in claim 1, wherein the cavities are distributed in substantially rotational symmetry to the axis of the rotor.

7. A cooling device according to claim 1 including a second condensation chamber on the other front face of said rotor, each of said cavities communicating with said second condensation chamber, said second chamber being included in said hermetically sealed enclosure in which said evaporable working medium is disposed, said second condensation chamber being defined at least in part by a second folded heat exchange surface having an area which is greater than the area of the front face of the rotor on which said second condensation chamber is disposed, said second heat exchange surface being defined by a plurality of substantially circular and concentrically positioned folds, the outside of said second heat exchange surface being exposed to a heat exchange medium to cool the working medium in said hermetically sealed enclosure, the inside of said second heat exchange surface providing a condensation zone for said working medium, whereby centrifugal force is utilized for providing return flow of the condensed working medium from said second condensing chamber to said plurality of cavities at the periphery of said rotor where said working medium is evaporated.

9. A cooling device as claimed in claim 8, wherein the bellows folds are coaxial and concentric with the axis of the rotor and the opening angle 2 gamma of the bellows folds complies with the equation