Background and Summary

5 The invention arose during further development efforts regarding marine trolling motors as shown in U.S. Patents 4,009,677, 4,099,478 and 4,305,012, incorporated herein by reference.

In the art of marine trolling propulsion, a

10 small electric driven outboard motor unit is mounted to the bow or transom of a boat and employed for slow speed maneuverability and positioning for fishing. The electric drive motor is housed within a lower submerged housing unit and coupled to drive a suitaole trolling

15 propeller. A separate high power internal combustion outboard motor is normally provided for rapid propulsion of the boat to the fishing location, after which the trolling motor is employed. For trolling, thrust from the trolling propeller is first transferred

2.0 from the propeller to the electric trolling motor housing and then to the boat.

Electric trolling motors are often designed to operate from a conventional 12 volt battery. Higher thrust units operating from a 24 volt supply are 5 available, but require two 12 volt batteries. A direct drive connection of the electric motor to the propeller provides a simple and reliable trolling motor with minimal cost. However, direct drive is subject to an inherent trade-off in efficiency because the motor has 0 an optimum operating speed which is higher than that of the trolling propeller.

The above noted patents show planetary gear reduction systems for driving a large diameter low rpm trolling propeller from a higher rpm electric trolling 5 motor. The large slower turning trolling propeller is

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more efficient in producing thrust at low boat speeds than high speed propellers. The larger trolling propeller requires more torque to turn it, which torque is provided by the gear reduction between the motor and the propeller.

In the present invention, a marine trolling propulsion system is provided with a traction drive speed reducer between the motor and propeller. Gears and gear teeth noise are eliminated, and replaced by rolling contact.

The invention is particularly advantageous in various design features which facilitate low manufacturing cost. The design inherently accommodates tolerance deviations in those members providing the loading of the rolling contact members. There is particular tolerance of axial deviations affecting radial loading of traction rollers. There is also tolerance in various members for movement transverse to the axis of rotation. The design enables different axes of rotation of the propeller shaft and the electric motor drive shaft such that such shafts do not have to be exactly aligned. The design enables a floating intermediate shaft which self-seats and tracks the rollers, without being constrained to align with either the input drive shaft from the electric motor or the propeller shaft. Inner race structure is keyed to the floating intermediate shaft and is designed to rock about such shaft transversely to the axis of rotation to equalize contact loading and to enable the inner race structure to self-seat and take its alignment from the rollers. The elimination of the alignment requirements of the rotating shafts eliminates concerns of binding and the strict manufacturing tolerances otherwise required to prevent same, including the attendant high machining costs, etc.

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The design provides overload slip protection enabling the electric motor drive shaft to keep rotating even if the propeller is momentarily stopped, e.g. upon striking a stump, whereby to avoid mechanical breakage within the drive. In addition, uniform loading of the normal contact force on the rollers is provided by applying substantially constant axial force versus axial deflection. This reduces manufacturing cost, provides increased and uniform life by avoiding overly loaded rollers, provides uniform overload slip capability, and accommodates mechanical wear without a reduction of the radial loading of the rollers.

In the preferred embodiment, planet cage structure locates off a three roller ball pattern. The planet cage structure is also designed for limited rocking about a point on the propeller shaft center to ensure equal division of roller contact loading from propeller thrust, and to transmit propeller thrust through the traction rollers and outer race to the housing structure without a separate thrust bearing. In combination with above noted rocking inner race structure, this further provides uniform and equally distributed roller contact engagement.

Brief Description of the Drawings

FIG. 1 is a side elevation view of a marine trolling propulsion system.

FIG. 2 is a partial sectional view of the lower housing unit in FIG. 1.

FIG. 3 is a sectional view taken along line 3-3 of FIG. 2.

FIG. 4 is an exploded isometric view of a portion of the structure in FIG. 2.

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FIG. 5 is an enlarged partial sectional view of a portion of the structure in FIG. 2.

FIG. 6 is a sectional view taken along line 6-6 of FIG. 5. FIG. 7 is an enlarged partial sectional view of a portion of the structure in FIG. 2.

FIG. 8 is a partial sectional view taken along line 8-8 of FIG. 2.

FIG. 9 is a partial sectional view taken along line 9-9 ^* of FIG. 2.

FIG. 10 is a partial sectional view showing an alternate embodiment of a portion of FIG. 2.

FIG. 11 is a sectional view taken along line 11-11 of FIG. 10.

Detailed Description

FIG. 1 is similar to FIG. 1 of incorporated U.S. Patent 4,099,478, and shows a marine propulsion trolling system 2. Generally, a boat for bass fishing and the like includes a forward horizontal deck to which the trolling system is pivotally secured. The trolling system has a lower submerged housing unit 4 having an electric motor 5, FIG. 2, including motor housing 5a, rotary armature 5b coupled to drive trolling propeller 6, and permanent magnets 5c Lower housing unit 4 is secured to the lower end of a support tube member 7 which in turn is suitably rotatably journaled in swivel tube support 8 which terminates at the upper end within a steering head 9. Swivel tube 8 is connected to pivotal mount assembly 10. Motor 5 is selectively energized from within the boat to vary the speed of propeller 6 and thereby the propulsion force, and tube 7 is pivoted within swivel tube support 8 to turn the boat. Pivot assemblies and supports for

trolling motors are shown in U.S. Patents 3,999,500, 4,008,680 and -4 ,051,802, incorporated herein by reference.

Electric motor 5 drives rotary drive shaft 12 which in turn drives rotary intermediate floating shaft 14. Coupler 16 couples drive shaft 12 to intermediate floating shaft 14 and enables different axes of rotation of shafts 12 and 14 such that these shafts need not be exactly aligned. This reduces manufacturing cost and eliminates concerns with binding. Drive shaft 12 has a screw driver flat head tip configuration 18, FIGs. 4 and 5. Intermediate floating shaft 14 has a notched screw .driver flat head tip configuration 20. Coupler 16 is an outer cylindrical shell 22 having a cross-shaped aperture 24 with perpendicular crossed slots 26 and 28 therethrough, FIG. 6. Flat head tip 18 is. loosely received in cross slot 26, and flat head tip 20 is loosely received in cross slot 28. Central notch 30 of tip 20 receives flat head tip 18 therein, to reduce overall axial length. This overlap and reduction of axial length is shown in FIG. 5. Alternatively, tip 18 can be notched. Further alternatively, neither tip is notched if the extra axial length is not objectionable. Rotation of drive shaft 12 drives shell 22 to rotate the shell which in turn drives intermediate floating shaft 14. The looseness of flat head tips 18 and 20 in respective cross slots 26 and 28 permits movement transverse to the axes of rotation of shafts 12 and 14 such that shafts 12 and 14 need not be exactly aligned.

Shaft 14 has a pair of axially spaced raised shoulders 32 and 34, FIGs. 4 and 7, of increased outer diameter and of a keyed configuration such as hexagonal. Split inner race structure is provided by

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axially spaced race halves 36 and 38 around shaft 14 and having respective inner diameters 40 and 42 with hexagonal configurations keyed to respective shoulders 32 and 34 and driven thereby such that rocation- c3 shaft 14 drivingly rotates race halves Z - ana 38. Race

Salves 36 and 38 slide over and are received on εfeoulders 32 and 34 with respective clearance gaps 44 smά 46, FIG. 7, to enable rocking of race hal/es 36 and 3_S_ about shoulders 32 and 34 transversely to the axis σf: rotation of axial shaft 14. The limit of rocking is also determined by respective gaps 48 and 50 between respective inner diameters 40 and 42 of the race halves and shaft 14.

Inner race halves 36 and 38 have curved respective bearing surfaces 52 and 54 engaging planet roller means 56 provided by three rollers or balls 57, 58 and 59 equally spaced therearound. The three balls 57-59 are driven by rolling contact with inner race halves 36 and 38 on shaft 14. The driven balls roll against an outer race 60 which is concentric to the balls and is fixed to housing portion 62 by press fit. Housing portion 62 is partially telescopically received in motor housing portion 64 and fixed thereto by through bolts, one of which is shown at 63. The bolts also attach nose cone 65 to motor housing 5a.

The shorter the axial gap between inner race halves 36 and 38, the greater the radially outward loading of balls 57-59 against outer race 60. Axial loading means 66 is provided for applying an axial force along shaft 14 urging race halves 36 and 38 towards each other to in turn urge balls 57-59 radially outwardly against outer race 60. This axial loading means preferably has a limited maximum loading force capability to intentionally provide overload slip protection to enable shafts 12 and 14 to continue to

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rotate even if rotation of planet carrier cage 68, to be described, and rotary driven output propeller shaft 70 is stopped. The continued rotation is enabled by sliding rather than rolling contact cf balls 57-59 with - inner races 36 and 38 or outer race 60.

In the preferred embodiment, axial loading is provided by one or more Belleville spring washers 72, FIG. 7, around shaft 14 and bearing on one side against retainer washer 74 stopped against snap ring 76 in 0 annular groove 78 in shaft 14, and bearing on the other side against race half 36. The other race half 38 engages and is stopped against stamped frusto-conical washer 80 which is stopped against snap ring 82 in annular groove 84 in shaft 14. Washer 80 extends 5 partially radially away from shaft 14 and partially axially toward the race halves to engage an outer curved spherical surface 86 of race half 38 along a tangent 88 having a perpendicular force vector 90, dashed line, FIG. 7, intersecting the center 92 of 0 shaft 14 at shoulder 34, such that race half 38 rocks about shoulder 34 without causing axial movement of washer 80 and self-seats to balls 57-59 to provide uniform rolling contact force and accommodate tolerance deviations. Belleville spring washers are preferred because they can be proportioned to apply a substantially constant axial force over a range of axial deflection, which is a known property of such washers. Belleville spring washers permit some cocking of race half 36 to equalize contact load on the balls. The constant axial force over a range of axial deflection provided by the Belleville spring washers is desirable to accommodate axial tolerance deviations of the components and their location and the machining required for grooves such as 78 and 84, all without

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substantially altering the radial loading of balls 57- 59. This accommodation of axial tolerance deviations reduces manufacturing cost. This also provides increased and uniform life by avoiding overloading of balls.57-59. This alsc provi- -ss uniform overload slip S-ijpa lity_. This also accommodates mechanical wear σv-er the Life of the system without a reduction cf the radϊial lc.sdlinig af balls 57-59- Face halves 36 and 33 take their alignment from balls 57-59 and can ec about shcni&sEe s 32 and 34 accordingly. Shaft 14 takes its alignment from the position of race halves 36 and 38. Coupler 16 permits angular misalignment of shaft 14 relative to shaft 12. Such structure further permits axial nonregistry of shafts 14 and 12. Each of balls 57-59 engages each of inner race halves 36 and 38 along curved bearing surfaces 52 and 54 and slightly elastically deforms- over a given deformation length therealong, as shown at 94, FIG. 7. An angle of engagement 96 is defined by a first line 98 drawn from the center 100 of the ball to the center 102 of elastic deformation length 94, and a second line 104 extending from the center 100 of the ball radially perpendicularly relative to shaft 14. The angle of engagement 96 is preferably about 20°, though it can be in the range of about 5° to 30°. An angle of engagement greater than 30° is not preferred because of loss of speed reduction capability. An angle of engagement less than 5° is not preferred because dimensions and tolerances become too critical to manufacture. As the contact points on the inner race halves 36 and 38 come closer together with smaller angle of engagement, so do the hexagonal drive shoulders 32 and 34, which in turn reduces directional control of shaft 14 resulting in increased wobble of the axial ends of shaft 14.

The center of elastic deformation engagement 102 is substantially radially aligned with shoulder 32 on shaft 14. The center of elastic deformation of each of balls 57-59 with race half 38 is likewise radially aligned with shoulder 34. The three balls 57-59 engage the two inner race halves 36 and 38 at six contact points with equally distributed contact pressure among the six points.

Housing section 105 is threadingly mounted to housing section 62 and closes a cavity therein which is filled with traction fluid, such as outboard motor gear oil to provide elastohydrodynamic lubrication. Another traction fluid is that sold under the name Santotrac by Monsanto Industrial Chemicals ^' Company. It is preferred that the traction fluid include extreme pressure (EP) additives, "Lubricants For Traction Drives", Machine Design, May 2, 1974, page 7, Monsanto Industrial Chemicals Company. The EP additives are preferred to prevent galling of the rolling contact surfaces when sliding due to overload. In the preferred embodiment, intentional overload slip capability is provided.

Planet cage 68 has three fingers 106, 107 and 108 extending axially between adjacent respective balls 57 and 58, 58 and 59, and 59 and 57. Each finger has sides which are curved, as shown at 109, to approximately the same curvature as the roller ball and which engage a respective roller ball. Planet cage 68 is located and positioned by balls 57-59, i.e. cage 68 takes its location from balls 57-59. As balls 57-59 roll along the inner and outer races, the balls slidingly engage and drive the sides of the cage fingers 106-108 to rotate cage 68 about shaft 14. Fingers 106-108 are joined by a central hub 110 having three abutment surfaces 111, 112 and 113 between adjacent respective fingers and each engaging a respective ball 57-5S for

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transmitting thrust. Hub 110 is frusto-conically tapered and then radially enlarged at axial end 114. Axial end 114 has an axial o ening 116, FIG. 2. with a radial slot 118 extending radially outwardly from axial opening 116. Axial end 120 of propeller shaft 70 has a radial aperture 122 therethrough through which extends a transverse pin 124 extending radially from axial shaft 70 and- loosely into radial slot 118 in cage 68. Rotation of cage 68' drives pin 124 and propeller shaft 70. Cage 68 acts as a transfer member to transfer rotation of shaft 14 and balls 57-59 to rotation of propeller shaft 70. The looseness of pin 124 in radial slot 118 permits rocking of cage 68 about center 123 of propeller shaft 70 at pin 124 such that cage 68 and propeller shaft 70 can be angularly misaligned.

Axial opening 116 in cage 68 has an inner rounded spherical abutment surface 126. Axial end 120 of propeller shaft 70 has a rounded spherical end face 128 engaging rounded abutment surface 126 such that cage 68 may rock about the center 123. Preferably the clearance between the balls 57-59 and cage 68 is small enough to limit the rocking of cage 68, to eliminate any potential binding between the parts, while providing running clearance between the parts. It is preferred that the cage movement be limited by the balls when at rest or in reverse. In other alternate embodiments, planet cage structure 68 can be loosely piloted on the inner diameter of outer race 60, or on the outer diameter of inner race 36, 38 or can be fixed to propeller shaft 70. In any event, during forward operation the clearance between the piloting surfaces will eliminate unnecessary friction while preventing excess rocking of cage 68 when the device is not running or is in reverse. Propeller thrust is transmitted to cage 68 at thrust surface 126 and in

turn transmitted by abutment surfaces 111-113 to balls 57-59 near the axis of rotation to minimize friction, and to outer race 60, without a separate thrust bearing for absorbing such propeller thrust. This provides an efficient forward thrust bearing without the cost of a separate antifriction bearing. Thrust in the reverse direction is absorbed by pin 124 against thrust washer bearing 130 in housing section 105. Housing section 105 includes a bushing 134 for supporting propeller shaft 70 and absorbing side shock loads on the propeller shaft, and is sealed by lip seal 135. Propeller 6 is mounted at hub 136 to propeller shaft 70 by nut 138. Hub 136 is supported directly on shaft 70. Hub 136 has a radial slot 142, FIGs. 2 and 9. Shaft 70 has a radial aperture 144 therethrough through which extends a transverse pin 146 extending radially from axial shaft 70 into radial slot 142 in hub 136. Rotation of shaft 70 drives pin 146 and hub 136 to rotate propeller 6. In the preferred embodiment, the invention enables the use of a propeller without a rubber insert in the hub. In typical prior applications, a rubber hub is inserted into a bore in the propeller. The outer diameter of the rubber hub insert frictionally drives the propeller, but will deflect or slip at a given overload, to protect the propeller if it should strike a stump or the like. However, the stiffness of the rubber insert varies, which in turn varies contact pressure, and hence the torque at which slip will occur. Furthermore, as the rubber ages, it sometimes bonds to the material of the propeller, such as aluminum, and won't slip. The present invention enables elimination of these problems by eliminating the need for a rubber hub insert.

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In an alternate embodiment, FIGs. 10 and 11, outer race 60a is fixed to transfer member 68a to rotate therewith. Transfer member 68a drives pin 124a and propeller shaft 70a. Stops in the form of bolts 151, 152 and 153, FIG. 11, are threaded into housing 62a and provide a cage holding balls 57a-59a against orbiting about central intermediate floating axial shaft 14a. Balls 57a-59a are in rolling contact with inner races 36a and 38a and outer race 60a, but do not orbit about shaft 14a. Bolts 151-153 are provided with respective roller bearings 154, 155 and 156 therearound for engagement with the rolling balls 57a-59a, allowing the balls to spin about respective spin axes parallel to axial shaft 14a. Housing portion 62a includes three integral thrust pads, one of which is shown at 162, FIG. 10, for engaging and stopping respective balls 57a-59a and absorbing the thrust from the center of the ball. The embodiment in FIGs. 10 and 11 reverses the direction of rotation. In another alternate embodiment, axles replace fingers 106-108, and roller bearings or needle bearings are provided therearound to further reduce friction, though such embodiment does increase cost. In another alternative, variable axial loading clamping force may be applied to urge inner race halves 36 and 38 towards each other in proportion to transmitted torque, for example by a screw or cam which provides increasing axial force with increasing motor torque. Overload slip protection may be provided in this embodiment with a preloaded spring that will deflect to limit the maximum applied load.

Roller balls 57-59 are preferably a steel ball bearing type material, hardened to 60 Rockwell C or higher. Races 36, 38 and 60 should likewise be bearing quality steel of equal hardness. Cage 68 is

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plastic of suitable strength and should have some bearing qualities for roller ball engagement, for example nylon.

It is recognized that various equivalents, alternatives and modifications are possible within the scope of the appended claims.

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