GB2110797A

GB2110797A - Temperature-controlled fluid coupling

Info

Publication number: GB2110797A
Application number: GB08234528A
Authority: GB
Inventors: Gerd Brunken
Original assignee: Fichtel and Sachs AG
Current assignee: ZF Sachs AG
Priority date: 1981-12-05
Filing date: 1982-12-03
Publication date: 1983-06-22
Also published as: GB2110797B; FR2517775A1; JPS58106230A; DE3148182A1; FR2517775B1

Abstract

The temperature-controlled fluid coupling for the fan of an internal combustion engine comprises two coupling parts (3, 7) rotatable in relation to one another about a common axis (5) of rotation, while a working chamber of a first coupling part (7) carrying fan blades (9) encloses the second coupling part (3), which is driven by the internal combustion engine. Between mutually adjacent faces of the working chamber (13) and of the second coupling part (3) several shear gaps (25, 27), filled with shear fluid in the operating condition, for the transmission of the drive torque extend between different radii. The width of the shear gaps (25) is dimensioned in dependence upon the radial distance from the axis of rotation, and increases with increasing distance. In this way a more uniform shear loading of the shear fluid can be achieved. The disengagement time of the fluid coupling is reduced. <IMAGE>

Description

SPECIFICATION Temperature-controlled fluid coupling The invention relates to a temperature-controlled fluid coupling for a fan of an engine, especially an internal combustion engine.

A temperature-controlled fluid coupling or liquid friction coupling for the fan of an internal combustion engine is known from German Publication Specification No. 2,814,608. The fluid coupling comprises a rotor of somewhat disc form driven in rotation by the internal combustion engine and a housing which encloses the rotor, is rotatable in relation to the rotor and carries the fan blades. The axial surfaces of the rotor, with adjacent faces of the housing working chamber enclosing the rotor, form substantially radially extending shear gaps which in the operating condition are filled with a shear fluid which transmits the drive torque from the rotor to the housing.

The shear fluid is a liquid of high viscosity.

Axially beside the working chamber a reservoir space is provided in the housing for the shear fluid, and is connected with the shear gaps by way of a feed opening. The shear liquid can enter the working chamber from the reservoir space by way of the feed opening. In the radially outer region of the shear gaps there is provided a pump device, formed for example by a displacement body, which conveys the shear fluid back out of the working chamber into the reservoir space. The feed opening comprises a temperature-controllable shut-off valve which is opened when the temperature rises and is closed when the temperature drops, for example by a bimetallic strip.

When the shut-off valve is closed the pump device delivers the shear fluid out of the working chamber, whereby the fluid coupling is uncoupled and the fan stops.

The torque transmitted by the shear fluid depends upon the width of the shear gap. The larger is the width, the lower is the transmitted torque, other conditions being equal. In order to achieve a defined operating behaviour of the fluid coupling, the width of the shear gaps must be maintained relatively precisely. This requirement presumes narrow manufacturing tolerances, having an unfavourable effect upon the production costs of the fluid coupling. In the known fluid coupling as explained above the rotor is formed as a disc which is flexible in the axial direction, is centred by the pressure of the shear fluid towards the middle of the working chamber, and ensures gaps of substantially equal widths.

In the known fluid coupling a compromise must be reached between the flexibility of the rotor and the transmittable torque. Since in the known fluid coupling the shear gaps have constant width in the radial direction, the shear fluid is stressed differently in dependence upon the radius. The shear rate and thus the stressing of the shear fluid are in proportion to the local difference of circumferential speed between the housing and the rotor and further in inverse proportion of the width of the shear gap at the location of the circumferential differential speed. The circumferential differential speed increases with increasing radius, so that in the case of shear gaps of equal width everywhere the loading of the shear fluid increases radially outwards.In the radially outer region of the shear gaps in the extreme case unacceptably high shear forces and unacceptably high temperatures of the shear fluid can occur, which in the short term influence the torque transmission and in the long term destroy the shear fluid.

Furthermore the idling behaviour of the known fluid coupling is unfavourable, since the pump device arranged in the radially outer region of the shear gaps cannot completely empty the shear gaps by reason of the counter-pressure of the shear fluid in the reservoir space caused by centrifugal force. Thus the known fluid coupling has a relatively high residual torque.

A further difficulty of known fluid couplings consists in that completely filled shear gaps can be emptied only relatively slowly after the closing of the shut-off valve. When the shear gaps are completely filled with shear fluid the pumping effect of the pump device, by reason of the still comparatively low torque difference between rotor and housing, is likewise low.

Thus the emptying of the shear gaps commences with delay. Initially the transmitted torque decreases only slightly, since at first the emptying, radially inner regions of the shear gaps contribute only slightly to the transmitted torque, in conventional fluid couplings.

It is the problem of the invention to avoid the above-explained disadvantages of conventional fluid couplings. A fluid coupling for the fan, especially of an internal combustion engine, is to be produced which is simple in construction and in which the shear fluid at the same time is stressed more regularly.

Furthermore the stopping delay is to be shortened and the torque transmitted in the disconnected condition (idling torque) is to be reduced.

In the solution to this problem the invention starts from a temperature-controlled fluid coupling for a fan, especially of an internal combustion engine, having the following features: a) Two coupling parts rotatable in relation to one another about a common axis of rotation, b) A working chamber in a first one of the two coupling parts which encloses the second of the two coupling parts, c) At least one shear gap extending between different radii and/or several shear gaps, extending on different radii, between mutually adjacent faces of the working chamber and of the second coupling part, d) a reservoir space for a shear fluid arranged axially beside the working chamber in the first coupling part, e) At least one feed opening for the shear fluid leading from the reservoir space to the shear gap or gaps, f) A pump device on the radially outer side of the feed opening, pumping the shear fluid out of the shear gap or gaps into the reservoir space, and g) A temperature-dependently controllable valve for the control of the shear fluid flow through the feed opening or the pump device.

Now the improvement consists in that the width of the shear gap or gaps is dimensioned in dependence upon the radial distance of the shear gap or gaps from the axis of rotation and increases with increasing distance. By reason of the increase of width of the shear gaps with increasing distance from the rotation axis, a more uniform distribution of the circumferential differential speed within the shear gaps and thus a more uniform shear stressing of the shear fluid result. Furthermore the object is achieved that the radially inwardly placed shear gaps or regions of the shear gaps produce a greater contribution to the transmitted torque, so that the pumping effect on disconnection of the fluid coupling commences more rapidly. The more rapidly commencing pump action is of importance especially for the cold-start behaviour of the fan coupling.When the internal combustion engine is started cold, the shear gaps are extensively filled with shear fluid, by reason of the preceding standstill time, and must be emptied as quickly as possible in order to avoid further cooling of the engine.

The locally transmitted specific torque depends, apart from the values specific to the fluid, especially the viscosity of the shear fluid, upon the radius in the fourth power, and is in inverse proportion to the local shear gap width. In radially outer regions of the shear gaps the gap width can be increased, whereby the residual torque transmitted by the disconnected fluid coupling, and thus the idling rotation rate of the fluid coupling, are reduced. The more uniform shear loading of the shear fluid in the shear gaps avoids overstressing of the shear fluid, especially in the radially outwardly situated regions. Thus the life of the shear fluid is prolonged. On the other hand the fluid coupling can be of more compact construction, since the shear fluid is subjected to no local peak loadings which otherwise would limit the maximum transmittabie torque.

In a preferred form of embodiment the width of the shear gap or gaps increases substantially in direct proportion to the radial distance from the axis of rotation. The increase is expediently limited to the radial region of the shear gaps between the feed opening and the pump device. Due to these measures it is possible to achieve an especially uniform shear stressing of the shear fluid. Substantially radially proceeding shear gaps are dimensioned so that they widen constantly linearly. Such shear gaps can be produced especially simply.

Especially in simple versions of the fluid coupling a tilting play which cannot be neglected occurs between the two coupling parts, and greatly influences the width of the shear gaps and thus the transmitted torque.

Disadvantages of this kind arise especially in fluid couplings where the shear gaps extend substantially radially. The width of the radially outer regions of such fluid couplings can vary relatively greatly by reason of the tilting play.

The outer regions specifically of the shear gaps supply the largest contribution to the transmitted torque, and are expecially sensitive to variations of width. This effect too is reduced by the invention, since the radially inner regions of the shear gaps, the variation of width of which is less than that of the radially outer regions by reason of the smaller shear gap width, are utilised to an increased extent for the torque transmission. In a preferred form of embodiment it is provided that the means width of those shear gap regions which are placed radially within the shear gap radius with the greatest variations of width caused by the tilting play, is smaller than the mean width on this shear gap radius.The mean width of the shear gaps on the radius with the greatest variation of width is expediently dimensioned approximately equal to or a little greater than this maximum variation of width, in order to dimension the fluid coupling for the largest possible torque to be transmitted.

In fluid couplings with substantially flat rotor of disc form it is preferably provide that the rotor has a constant thickness in the region of its gap-forming faces, while the gapforming faces of the working chamber are suitably dimensioned in order to achieve radially outwardly widening shear gaps. In this way it is possible to form the rotor as a simple punching while the form of the gap-forming face of the partition between working chamber and reservoir space can be stamped.

Examples of embodiment of the invention are to be explained in greater detail below with reference to the drawings, wherein: Figure 1 shows a diagrammatic sectional view of a temperature-controlled fluid coupling with substantially radially extending shear gaps; Figure 2 shows a fluid coupling similar to that according to Fig. 1, differing as regards the configuration of its shear gaps, and Figure 3 shows another form of embodiment of a fluid coupling for the fan of an internal combustion engine, having a potshaped rotor.

The fan coupling according to Fig. 1 comprises a rotor 3 seated fast on a shaft 1, formed as a substantially flat circular disc and rotating together with the shaft 1 about an axis 5 of rotation. A coupling housing 7 is mounted freely rotatably in relation to the rotor 3 on the shaft 1, in a manner not further illustrated. In this case the shaft 1 is sealed off in relation to the housing 7. The housing 7 carries several radially protruding fan blades 9 offset in relation to one another in the circumferential direction. A partition 11 separates the interior of the housing 7 into a working chamber 1 3 enclosing the rotor 3 and a reservoir space 15, arranged axially beside the working chamber 13, for a viscous shear fluid.The rotor 3 has substantially axially extending lateral faces 17, 1 9 which with opposite wall faces 21 and 23 of the working chamber 1 3 form two substantially radially extending shear gaps 25 and 27 respectively.

During fan operation the shear gaps 15, 1 7 are filled with shear fluid, whereby the rotor 3 exerts a torque upon the housing 7, by way of the shear forces transmitted by the shear fluid.

The fluid coupling is capable of coupling and uncoupling in dependence upon temperature. For this purpose a feed opening 33 is provided in the partition 11 of the radius of the filling level, entered at 31, of the disengaged coupling, which opening is closable by a shut-off valve, for example in the form of a leaf spring 35. A control member 37 opens the feed opening 33 for the shear fluid on exceeding of a predetermined temperature level and closes it on falling short of the temperature level. The control member 37 can be a bimetallic strip or an electrically controllable drive system. The feed opening 33 opens into a radially inwardly placed region of the working chamber 13, and thus of the shear gaps 25, 27, which may be connected by passages 38 in the rotor 3.In the radially outwardly situated region of the working chamber 13, close to the external circumference of the rotor 3, a return opening 39 is provided in the partition 11 and together with a displacement body 41 forms a pump device which delivers the shear fluid from the working chamber 1 3 back into the reservoir space 1 5.

When the fluid coupling is closed the shutoff valve of the feed opening 33 is opened, whereby the shear fluid from the reservoir space 1 5 can enter the shear gaps 25, 27.

The feed opening 33 is dimensioned so that more shear fluid can enter the working chamber 1 3 than flows back through the return opening 39. The working gaps 25, 27 are filled with shear fluid and transmit the operating torque from the rotor 3 to the housing 7.

After the feed opening 33 is closed the displacement body 41 pumps the shear fluid back out of the working chamber 1 3 into the reservoir space 15, until the coupling is disengaged.

The shear gaps 25, 27 widen linearly with increasing radius, starting from an inner radius r to the circumferential radius ra. The inner radius r; is equal to the distance of the centre point of the feed opening 33 from the axis 5 of rotation. The other raid us era reaches substantially to the external circumference of the rotor 3. The local shear rate and thus the shear stressing of the shear fluid are dependent upon the quotient of the differential circumferential speed of the rotor 3 and the housing 7 to the width of the shear gap at the same radius.

If as explained above the width t of the shear gaps 25, 27 increases in direct proportion to the radius r, then the ratio r t and thus the shear rate remains constant within the limits of the radii rj and ra, namely at a value which is determined by the ratios ri:ti= r,:t,.

Here tj is the gap width at the radius r and ta is the gap width at the radius ra. The shear fluid is therefore uniformly stressed on every radius within the shear gaps 25, 27.

In conventional fluid couplings the internal width of the shear gap is made equal to the external width, so that the shear rate at the external circumference is greater than at the internal circumference of each shear gap. The constant width of the shear gap, in conventional fluid couplings, is so dimensioned that the shear fluid is adequate for the shear loading at the external circumference. The width ta of the fluid coupling according to the invention is dimensioned according to the same consideration. On the other hand however towards the rotation axis 5 the width of the shear gap decreases, from which essential advantages result.

In the radially inner region of the shear gaps a greater torque is transmitted than in conventional fluid couplings with constant gap width. As a whole therefore with fluid couplings according to the invention it is possible for greater torques to be transmitted or, in the case of equal torque, the coupling can be of more compact dimensions.

Furthermore the delay in disengagement of the coupling is reduced. Due to the tapering of the radially inner region of the shear gaps, the shear gap volume is reduced and the pump-away time is shortened. In the pumping away, firstly the radially inwardly situated region of the shear gaps begins to empty. Since the proportion of the radially inwardly placed regions of the shear gap in the torque transmission was increased in comparison with conventional fluid couplings, in the disconnection of the fluid coupling the torque initially decreases more rapidly than in conventional couplings.By reason of the smaller gap width in the radially inwardly situated region the radial width at of the shear fluid volume AV pumped out of the shear gap per unit of time is greater in the coupling according to the invention that in conventional fluid couplings. Finally in the fluid coupling according to the invention the centrifugal pressure of the shear fluid, generated by the shear fluid in the reservoir space 1 5 and acting against the fluid flow through the return opening 39, is less, since the reservoir volume of shear fluid can be made smaller, as a result of the smaller volume of the shear gaps.

The housing 7 can be mounted with a certain tilting play in relation to the shaft 1 and the rotor 3. The width ta at the external circumference of the rotor 3 in this case is dimensioned so that even at the most unfavourable position of tilt of the rotor 3 in relation to the housing 7 the shear fluid is not unacceptably stressed. The variation of width of the shear gaps caused by the tilting play decreases towards the axis 5 of rotation. Despite the fact that the shear gaps 25, 27 narrow towards the axis 5 of rotation, the relative width variation caused by the tilting play remains constant at every point of the shear gap.

In a modification of the form of embodiment of the fluid coupling as explained above, the width ta of the shear gaps 25, 27 can also be dimensioned wider than in conventional fluid couplings, the consequent reduction of the torque contribution of the radially outer regions being compensated by the increase of the torque contribution of the radially inner regions. Due to the widening of the shear gaps in the radially outer region however the idling torque and thus the idling rotation rate of the fluid coupling are also reduced. This has advantages with regard to noise generation and the idling losses.

In the form of embodiment according to Fig. 1 the shear gaps 25, 27 widen along the entire radial range between the feed opening 33 and the return opening 39. The widening of the shear gaps can however also be limited to the radially inner region, while the radially outer region can be dimensioned according to other considerations, especially to improve the idling behaviour of the fluid coupling. By way of example in the radially outer region the gap width can be of constant dimensions or can have a rate of width variation differing from that of the radially inner region.

In the fluid coupling according to Fig. 1 the side faces 21 and 23 of the working chamber 1 3 extend parallel with one another and perpendicularly to the axis 5 of rotation. The side faces 1 7 and 1 9 of the rotor 3 are inclined conically towards one another. Fig. 2 shows another form of embodiment which differs from the form of embodiment according to Fig. 1 only in the configuration of the side faces. Therefore for the explanation of details of the fluid coupling according to Fig. 2, reference is made to the explanations referring to Fig. 1. Parts of like effect are here designated by reference numerals increased by the Fig. 100. The fluid coupling according to Fig.

2 has a rotor 103 formed as a shallow, substantially plane disc, arranged in a working chamber 11 3 of a housing 107. Between axial side faces 11 7 and 11 9 on the one part and adjacent axial side faces 121, 123 of the working chamber 11 3 on the other there are again provided substantially radially extending working gaps 125 and 127 which widen from radially inwards to radially outwards. In contrast to the fluid coupling according to Fig. 1 however the side faces 117, 11 9 of the rotor 103 extend parallel with one another, while the side faces 121 and 123 diverge in the radial direction. In this form of embodiment the rotor 103 can be punched at favourable cost out of a metal sheet of constant thickness.The side face 123 can be produced in a simple way by stamping of a partition 111, likewise formed as a shaped sheet metal piece, between the working chamber 11 3 and a reservoir space 11 5 for the shear fluid.

Fig. 3 shows another form of embodiment of a fluid coupling for the fan of an internal combustion engine. The fluid coupling differs from the fluid coupling according to Fig. 1 in the configuration of its rotor and the manner or arrangement of its shear gaps. For the explanation of the fluid coupling according to Fig. 3 reference is made to the description of the fluid coupling according to Fig. 1, parts of like effect being designated by reference numbers increased by the Fig. 200.

The fluid coupling comprises a rotor 203 of pot form secured on a shaft 201 and driven in rotation about an axis 205 of rotation. The rotor is enclosed by a housing 207 carrying fan blades 209. A partition 211 divides the interior of the housing 207 into a working chamber 213, which contains the rotor 203, and a reservoir space 215. A feed opening by way of which the shear fluid can enter the working chamber 213 from the reservoir space 215 is designated by 233. The feed opening 233 is controlled in temperature dependence by a valve 235. The shear fluid is conveyed back out of the working chamber 213 into the reservoir space 215 by a displacement body 241 by way of a return opening 239 in the partition.

The pot-shaped rotor 203 comprises a bottom part 251 extending perpendicularly of the axis 205 of rotation, from which part a hollow-cylinderical wall part 253 protrudes axially. The radially outer peripheral face 255 of the wall part 253, together with a radially opposite inner peripheral suface 257 of the housing 207, forms a shear gap 259 lying radially outside the rotor 203. The partition 211 in pot form follows the inner contour of the rotor 203 and has a radially outwardly directed peripheral surface 261 which lies radially opposite to the inner periphery 263 of the wall part 253 and forms a second, inwardly situated, shear gap 265. The shear gaps 259 and 265 extend in an axial direction. Each of the shear gaps 259 and 265 has constant radial width over its axial length.

However the shear gap 265 extending on a smaller radius has less radial width than the shear gap 259 extending on a larger radius.

The radial widths of the two shear gaps are dimensioned so that they have approximately equal shear rates.

The axial side faces of the bottom part 251 are at such great distance from the adjacent side faces of the housing 207 and the partition 211 that they do not contribute to the torque transmission. As indicated in Fig. 3 by chain lines 267 and 269, the distances can however be reduced so far that radially proceeding shear chambers are formed. These shear chambers are preferably dimensioned in accordance with the forms of embodiment of Fig. 1 or 2, and widen radially outwards.

In Fig. 3 the wall part 253 extends at right angles to the bottom part 251. The wall part 253 can also extend at a different angle, and especially be curved with steady curvature out of the bottom part.

In the forms of embodiment as explained above the shear gaps widen constantly linearly with increasing distance from the axis of rotation. Forms of embodinient in which the width of the shear gap increases suddenly with increasing radius, OF with width variation rate varying suddenly with increasing radius are likewise possible.

Claims

1. Ternperature-controlled fluid coupling for a fan, especially of an iiatepna8 combustion engine, comprising:a) two coupling parts (3, 7; 103, 107; 203, 207) rotatable in relation to one another about a common axis (5; 205) of rotation, b) a working chamber (13; 113; 213) in a first one Gf the two coupling parts (7; 1 07; 207), enclosing the second of the two coupling parts (3; 'it%; 203), c) at least one shear gap (25, 27; 125, 127) extending between different radii and/or s"v- eral shear gaps (259, 265) extending on different radii, between mutually adjacent faces of the working chamber and of the second coupling part, d) a reservoir space (15; 115; 215) for e shear fluid, arranged axially beside the working chamber in the first coupling part, e) at least one feed opening (33; 233) for the shear fluid extending from the reservoir space to the shear gap or gaps, f) a pump device (39, 41; 239, 241) on the radially outer side of the feed opening, pumping the shear fluid out of the shear gap or gaps into the reservoir space, and g) a temperature-dependently controllable valve (35; 235) for controlling the flow of shear fluid through the feed opening or the pump device, characterised in that the width of the shear gap or gaps is dimensioned in dependence upon the radial distance of the shear gap or gaps from the axis of rotation, and increases with increasing distance.

2. Fluid coupling according to Claim 1, characterised in that the second coupling part (3; 103) is formed as a disc and forms, in the region of each of its two axial side faces (17, 19; 117, 119), with adjacent axial faces (21, 23; 121, 123) of the working chamber (13; 113), a shear gap (25, 27; 1125, 1 27) which widens radially outwards.

3. Fluid coupling according to Claim 2, characterised in that the width of the shear gap or gaps (25, 27; 125, 127) increases radially outwards in the radial region between the feed opening (33) and the pump device (39, 41).

4. Fluid coupling according to Claim 2, characterised in that the disc (103) has subtantially constant thickness and in that the distance of the two faces (121, 123) of the working chamber (113), which form the shear gaps (125, 127), from one another increases radially outwards.

5. Fluid coupling according to Claim 1, characterised in that the second coupling part (203) is of pot form and forms, at least in the region of its external circumferential surface (255) and its internal circumferential surface (263), with adjacent faces (257; 261) of the working chamber (213), in each case an inner shear gap (265) end an outer shear gap (259), and in that the inner shear gap (265) has a smaller gap width than the outer shear gap (259).

6. Fluid coupling according to one of the preceding Claims, characterised in that the width of the shear gap or gaps (25, 27; 125, 127; 259, 265) increases substantially in direct proportion to the radial distance from the axis (5; 205) of rotation.

7. Fluid coupling according to Claim 6, characterised in that the width of the shear gap or gaps (25, 27; 125, 127) increases constantly with increasing distance from the axis (5) of rotation.

8. Fluid coupling according to Claim 1, wherein one of the two coupling parts (7; 1107) is mounted rotatably with limited tilting play on the other coupling part (1, 3; 103) and the shear gap or gaps (25, 27; 125, 127) extend between different radii, characterised in that the mean width of those shear gap regions which are placed radially within the shear gap radius (r") with the greatest width variation caused by the tilting play is less than the mean width of this shear gap radius (r,).

9. A Temperature-controlled fluid coupling for a fan substantially as described with reference to the accompanying drawings.