US3690196A

US3690196A - Geared drives incorporating fluid couplings

Info

Publication number: US3690196A
Application number: US142682A
Authority: US
Inventors: John Bilton
Original assignee: Fluidrive Engineering Co Ltd
Current assignee: Fluidrive Engineering Co Ltd
Priority date: 1970-05-13
Filing date: 1971-05-12
Publication date: 1972-09-12
Anticipated expiration: 1989-09-12
Also published as: DE2123481B2; DE2123481A1; DE2123481C3; GB1349147A

Abstract

A double-circuit fluid coupling has a rotating outer casing formed in two similar halves between which is clamped an internal flange of a gear rim formed with wide external gear teeth. Only the middle portions of the gear ring and casing are in contact while the axially outer portions are spaced by air gaps from the casing to form a much longer heat path from the coupling circuits to the gear rim and thereby prevent distortion of the gear rim as the result of uneven expansion due to sudden heat generation in the working circuits during maneuvring.

Description

United States Patent Bilton 1 Sept. 12, 1972 54] GEARED DRIVESINCORPORATING FLUID COUPLINGS [72] Inventor: John Bilton, Hampton, England [73] Assignee: Fluidrive Engineering Company Limited, Middlesex, England [22] Filed: May 12, 1971 [21] Appl. No.: 142,682

[30] Foreign Application Priority Data May 13, 1970 Great Britain ..23,240/70 [52] US. Cl. ..74/730, 60/54, 74/431, 74/665 B [51] Int. Cl ..F16h 47/06, F16d 33/00, F16h 57/04 [58] Field of Search .....74/730, 431, 432, 433; 60/54 [56] References Cited UNITED STATES PATENTS 2,298,310 10/1942 Ray ..60/54 2,465,919 3/1949 Novak ..60/54 2,549,557 4/ 1951 Yancho et a1. ..60/54 2,585,968 2/1952 Schneider ..60/54 x 2,588,668 3/1952 Syrovy ..60/54 FOREIGN PATENTS OR APPLICATIONS 1,157,849 7/1969 Great Britain ..60/54 Primary Examiner-Carlton R; Croyle Assistant Examiner-Thomas C. Perry AttorneyWoodhams, Blanchard & Flynn [57] ABSTRACT A double-circuit fluid coupling has a rotating outer casing formed in two similar halves between which is clamped an internal flange of a gear rim formed with wide external gear teeth, Only the middle portions of the gear ring and casing are in contact while the axially outer portions are spaced by air gaps from the casing to form a much longer heat path from the coupling circuits to the gear rim and thereby prevent distortion of the gear rim as the result of uneven expansion dueto sudden heat generation in the working circuits during maneuvring.

6 Claims, 6 Drawing Figures PATENTED EP 2 2 3.690.196

sum 2 [IF 5 MLZQM W nrfam PATENTEDSEP 12 I972 3.690.196

saw u I]? 5 GEARED DRIVES INCORPORATE G FLUID COUPLDIGS This invention relates to high-powered geared drives incorporating fluid couplings, for example marine propulsion drives transmitting say 25,000 h.p.

The pinions and gear wheels of such drives are necessarily wide (measured in the direction of the axis of rotation) in order to transmit the high power without excessive tooth loading. Moreover, the gear teeth must be very accurately formed in order to ensure substantially uniform load distribution over the gear teeth as they mesh in order to avoid localized overloading. It is found in practice that variations in the gear tooth profiles over quite small lengths of the teeth by fractions of a thousandth of an inch can cause large alterations in the local loading of the tooth and can therefore completely upset the uniform tooth loading which is desired along the full width of the gear wheel.

In addition to the width of the gear wheel itself, measured in the axial direction, further length is required to accommodate the supporting bearings for the gear wheel. The incorporation of fluid couplings into the drive in the conventional manner as shown for example in British Pat. specification No. 1,076,273 would obviously further increase the length of the fluid coupling and gear wheel assembly. In many cases, particularly in the case of warships, theoverall length of such a drive is undesirable or even unacceptable. In order-to provide a more compact drive, it has already been suggested that in the case of relatively small drives, at most of the order of a few hundred horsepower, a gear wheel rim should be secured to the rotating casing of a fluid coupling. An example of such an arrangement is shown in US. Pat. specification No. 1,979,930.

Problems arise however in applying this arrangement to high power drives.

In order to reduce the overall radial demensions of high power fluid couplings, it is known to use a double circuit coupling in which two interconnected. inner vaned elements, one foreach working circuit, face outwards away from each other and each form one element of the respective working circuit and the other two vaned elements face inwards and are interconnected by a casing enclosing the inner elements.

With this arrangement, it would appear that the casing could form the gear ring itself or that a separate toothed gear rim could be shrink-fitted onto it.

However, in accordance with the present invention, the gear wheel rim and easing are secured together only in the middle regions of the gear wheel rim and the casing in such a manner that the axial end portions of the gear wheel rim are spaced from the casing by an air This arrangement avoids the risk of heat generated in the coupling working circuits being directly conducted to one portion of the gear wheel very rapidly before other portions have the chance to reach equivalent temperatures.

Where the fluid couplings are selectively filled or emptied to select forward or reverse drive, abnormally high temperatures may be reached within the coupling over a short period during a crash change from forward astern drive or vice versa in a ship. Even though large quantities of liquid are circulated through the coupling for cooling purposes, for example 1,000 gallons per minute, the reversal of a 25,000 h.p. drive can generate so much heat that the temperature within the coupling may rise by about 230 F. to a final temperature of over 300 F.

With the arrangement of the invention, the air gaps between the two ends of the gear wheel rim and the outer casing of the coupling act as heat barriers so that the conduction path for heat from the outer portions of the coupling working circuits to the gear wheel rim is a relatively long path extending from an end of the casing to the middle and then upwards again through the gear wheel rim. By this means the rapid high rise in temperature is reduced and heat is transferred from the casing to the middle of the gear wheel rim at a lower rate such that it can be spread through virtually the whole of the gear wheel rim much more uniformly than if the whole of the gear wheel rim where in thermal contact with the casing.

The invention will now be described by way of example with reference to the accompanying drawings, in which:

FIG. 1 shows an axial sectional view of a combined double circuit fluid coupling and gear wheel rim incorporated in reduction gearing,

FIG. 2 shows a detail of 1 on an enlarged scale,

FIG. 3 is a diagrammatic end view of a. marine propulsion system incorporating such a combined fluid coupling and gear wheel assembly, 7

FIG. 4 is an elevational view seen in the direction of the arrows IV-IV of FIG. 3, and

FIGS. 5 and 6 are views similar to FIGS. 3 and 4 respectively of another marine propulsion system, in-

corporating the assembly shown in FIGS. 1 and 2.

The double circuit coupling and gear wheel assembly shown in FIGS. 1 and 2 comprises two vaned impeller elements 1 and 2 and two vaned runner elements 3 and 4 which are secured back-to-back. The elements 1 and 3 together define a first toroidal working circuit W while the elements 2 and 4 define a second toroidal working circuit W Suitable means 8,, S are provided for supplying working liquid to the working circuits W and W so as to maintain them filled when drive is to be transmitted from the impellers l and.2 to the runners 3 and 4. Provision is also made for emptying the working circuits when the drive is to be discontinued and for maintaining a flow of the working liquid for cooling purposes while the drive is being transmitted. These results may be achieved by conventional methods.

The runners l and 2 form the ends of a casing 6 for the coupling which includes two identical generally cylindrical portions 7 and 8 bolted to flanges 9 and 10 on the impellers l and 2 at 11 and 12 respectively. The axially inner ends of the casing portions 7 and 8 have inwardly directed flanges 13 and 114 between which is clamped an inwardly-directed central flange 15 of a symmetrical gear wheel rim 16 formed with long helical gear teeth 17. The flanges 13, 14 and 15 are clamped tightly together by a ring of bolts 18.

Adjacent the root of the flange 15, the gear wheel rim 16 has two cylindrical seating surfaces 19 and 20 which engage corresponding clearance spigot surfaces on the casing portions 7 and 8 to provide initial support for the gear wheel rim 16 during assembly. After the gear wheel rim 16 has been accurately positioned, holes are reamed through the flanges I3, 14 and 15 to accommodate dowel pins (not shown). As is clearly shown in FIG. 1, the remainder of the length of the gear wheel rim 16 is spaced from the adjacent casing portions 7 and 8 by air gaps 21 and 22. As a result, the heat conduction path from the working circuits W and W to the gear wheel rim 16 is long. Supposing for example that the working circuit W and W, are empty and that the impellers 1 and 2 are rotating in the opposite direction to that of the runners 3 and 4. If the working circuit W and W are suddenly filled, a large amount of heat will be rapidly generated in the working circuits and as a result arresting the machinery components attached to the runners 3 and 4, including the propeller shaft and propeller, and in accelerating them in the direction of rotation of the impellers 1 and 2.

As mentioned above, the temperature rise may be as high as 230 F. If the gear wheel rim 16 were indirect thermal contact with the casing portions 7 and 8 along the whole length of the gear wheel rim, the rapid temperature rise would be conducted first of all to the two ends of the gear wheel rim, at which stage the two end portions would begin to expand while the relatively cool center portions of the gear wheel rim retain their normal dimensions. In this way, the very accurately formed helical tooth profile would be distorted and the entire tooth loading would be concentrated over the high spots and thus over a relatively small proportion of the total tooth length. Serious damage would be inevitable.

fith the arrangement shown in FIGS. 1 and 2, the rapid temperature rise from the working circuits cannot reach any of the gear wheel rim 16 until it has been conducted through the casing portions 7 and 8 to the flange 15. Accordingly, the rate of rise of temperature at the flange will be less than at the outer ends of the casing portion 7 and 8. Moreover, the temperature rise is applied to the gear wheel rim over a relatively large area and at a relatively thick portion of the gear wheel rim so that the heat conduction path to all parts of the gear wheel rim is comparatively short. Thus the rate of increase in temperature is reduced by the casing portion 7 and 8 and is well spread over the whole length of the gear wheel rim.

. The underside of the gear wheel rim 16 where it forms one surface of the

air gap

31 and 32 may be profiled to control the rigidity of the gear wheel rim in such a manner that satisfactory tooth contact is achieved over the full length of the gear wheel rim in service. This profile may be initially calculated and checked by initial trials when the actual degree of tooth bedding can be visually checked. Thereafter any slight modifications to the profile can be made.

The marine power transmission shown in FIGS. 3 and 4 comprises an output bull wheel 31 secured to the propeller shaft 32. The hull wheel 31 is to be driven selectively in forward or reverse direction from two

unidirectional turbine shafts

33A and 33B. The shaft 33A carries a pinion 34A which meshes with an ahead wheel 35A and with the gear wheel rim of a fluid coupling and gear wheel assembly 36A of the kind shown in FIGS. 1 and 2. The coupling and gear wheel assembly 36A also meshes with an astem coupling and gear wheel assembly 37A of similar construction.

The ahead wheel 35A is connectable with a shaft 38A by means of a dog clutch 39A. The shaft 38A carries a pinion 40A which meshes with the bull wheel 31. A shaft 41A carries a pinion 42A which also meshes with the bull wheel 31. The shaft 41A extends through the ahead coupling and gear wheel assembly to carry a synchronous self-shifting clutch 43A having a control member 44A. The other part of the S.S.S. clutch 43A is secured to a hollow shaft 45A which in turn is secured to the two runner elements of the ahead assembly 36A. The moving elements of the dog clutch 39A and the control sleeve 44A of the SSS. clutch 43A are linked by a lever 46A pivoted at 47A. With this arrangement, the dog clutch 39A cannot attempt to engage until the SSS. clutch detects synchronism between the

shafts

45A and 41A during filling of the working circuit of the ahead assembly 36A.

In operation, ahead drive is engaged by emptying the working circuit of the astem coupling and gear wheel assembly 37A and thereafter filling the working circuits of the ahead assembly 36A until the

clutches

43A and 39A can be engaged. The power flow from the turbine shaft 33A to the bull wheel 31 is then shared by the two shafts 41A and 38A.

To engage astem drive, the speed of the turbine shaft 33A is temporarily reduced to unload the

clutches

39A and 43A. At the same time the working circuits of the ahead assembly 36A are emptied and the working circuits of the astem assembly 37A are filled. The gear wheel rim of the ahead assembly 36A then acts as an idler wheel to transmit power from the working shaft 33A to the astem assembly 37A, the output shaft 48A of which carries a pinion 49A meshing with the bull wheel 31.

In the marine propulsion system shown in FIGS. 5 and 6, each turbine shaft is coupled to an

input shaft

51, 52 carrying a

pinion

53, 54 respectively. The pinion 53-meshes with'the gear wheel rims of two ahead gear wheel and

coupling assemblies

55 and 56 of the kind shown in FIGS. 1 and 2, the assembly 55 being that shown in FIGS. 1 and 2. The pinion 54 meshes with two further coupling and

gear wheel assemblies

57 and 58. The output of each of the four coupling and

gear wheel assemblies

55, 56, 57 and 58 is connected to a

respective pinion

60, 61, 62, 63 all of which mesh with a common bull wheel 64.

A fifth coupling and gear wheel assembly 65 meshes with the two

assemblies

56 and 57 to provide an astem drive through a pinion 66 when its working circuits are filled and thus all the four ahead

couplings

55, 56, 57 and 58 are emptied. The astem pinion 66 also meshes with the bull wheel 64.

Where the turbines are gas turbines, the

input shafts

51 and 52 may rotate at 5,660 rpm and the gear wheel rims of each of the four assemblies 55 to 58 may then be arranged to rotate at 1,560 rpm while the astem gear wheel rim may rotate at 1,600 rpm. If each gas turbine produces 25,000 h.p. each ahead coupling and gear wheel assembly will transmit 12,500 h.p. and may be designed to be of a size such that the minimum slip under normal conditions is no more than 1% percent. Preferably, an over-running or disconnectible clutch is provided between each turbine and the

respective pinion

53 and 54 to enable one turbine to run while the other is stationary, for example for repairs or maintenance.

It will be appreciated that the gear wheel rim of the astern assembly 65 interconnects the two

input pinions

53 and 54 through the gear wheel rims of the

assemblies

56 and 57. Thus, when one turbine is out of service the drive from the other turbine is distributed over all four ahead coupling assemblies. Even if the two fluid coupling assemblies associated with the idle turbine are emptied the slip in the two remaining coupling assemblies would still be only about 2% percent.

While the temperature rise in the ahead couplings during maneouvring is reduced by the using of four ahead couplings, the temperature rise in the single astern coupling could still be serious if the astern assembly did not incorporate the features of the invention. Moreover, in order to reduce the total oil flow demanded from the filling pump system durint filling of the ahead couplings in a change to forward drive, the filling system may be arranged so that one fluid coupling assembly (e.g., 55) fills substantially before the other fluid coupling assembly (e.g., 56) associated with the same turbine. The application of the features of the invention to the ahead coupling assembles avoids the overheating problem during such a change to ahead drive. At the same time, by more quickly filling one coupling, the overall time required to effect a change in the direction of drive is reduced.

As can be seen from FIGS. 2 and 3 and from FIGS. 4 and 5, the use of the combined fluid coupling and gear wheel assemblies enables the overall length of the transmission and hence the size of its casing to be materially reduced.

I claim:

ii. In combination, a double-circuit fluid coupling and a gear ring, each circuit of said fluid coupling being defined by a first vaned member and a second vaned member, the two first vaned members being secured in back-to-back relationship, with casing means interconnecting the two second vaned members and enclosing the two first vaned members, wherein the gear ring is secured to said casing means only in the middle regions respectively of said gear ring and said casing so as to define air spaces between said casing means and the axially outer portions of said gear ring.

2. The combination set forth in claim 1, in which the axially outer portions of said gear ring are of progressively reduced radial thickness with decreasing distance from the ends of said gear ring.

3. The combination set forth in claim 2, in which the external surface portions of said casing means underlying the axially outer portions of said gear ring are correspondingly shaped so that said air spaces are annular and substantially parallel-sided.

4. The combination set forth in claim 1 in which said gear ring, said casing means and the two working circuits of said coupling are symmetrically disposed about a common median plane of the coupling and gear ring.

5. The combination set forth in claim 4, in which said I gear ring has a central internal flange and said flange is clamped between substantially identical casing portions of said casing means.

6. The combination set forth in claim 2, in which said gear ring, said casing means and .the two workingcircuits of said coupling are symmetrically disposed about a common median plane of the coupling and gear ring.

Claims

1. In combination, a double-circuit fluid coupling and a gear ring, each circuit of said fluid coupling being defined by a first vaned member and a second vaned member, the two first vaned members being secured in back-to-back relationship, with casing means interconnecting the two second vaned members and enclosing the two first vaned members, wherein the gear ring is secured to said casing means only in the middle regions respectively of said gear ring and said casing so as to define air spaces between said casing means and the axially outer portions of said gear ring.

5. The combination set forth in claim 4, in which said gear ring has a central internal flange and said flange is clamped between substantially identical casing portions of said casing means.

6. The combination set forth in claim 2, in which said gear ring, said casing means and the two workingcircuits of said coupling are symmetrically disposed about a common median plane of the coupling and gear ring.