Patent Abstract:
A two-speed transmission having an input shaft and an output shaft, the transmission being capable of transitioning between fixed ratios, the high-range ratio being direct 1:1 and the low-range ratio being about 2:1. The transmission is a simple lightweight, yet robust, configuration utilizing only two gear meshes, being comprised of an input gear, a cluster gear, and an output gear. The transmission is controlled with a clutch and a sprag and with the input and output shafts turning in the same direction.

Full Description:
ORIGIN OF INVENTION 
     The invention described herein was made by an employee of the United States Government and may be manufactured and used by or for the Government for Government purposes without the payment of any royalties thereon or therefore. 
    
    
     CROSS-REFERENCE TO RELATED APPLICATIONS 
     This application claims the benefit of U.S. Non-Provisional patent application Ser. No. 12/431,456 filed Apr. 28, 2009, entitled “Offset Compound Gear Inline Two-Speed Drive”, which issued as U.S. Pat. No. 8,091,445 B1 on Jan. 10, 2012. The entirety of the above-noted application is incorporated by reference herein. 
     TECHNICAL FIELD 
     The invention relates to transmissions, and more particularly to a device(s) and configurations which provide a simple, lightweight two-speed drive which can be used either as an overall transmission or as a supplemental add-on input transmission (e.g., over-drive/under-drive) to extend the capability of an existing transmission. 
     BACKGROUND 
     In several recent studies and on-going developments for advanced rotorcraft, the need for variable or multi-speed capable rotors has been raised. A speed change of up to 50% has been proposed for future rotorcraft to improve overall vehicle performance. Accomplishing rotor speed changes during operation requires both a rotor that can perform effectively over the operational speed-load range, and a propulsion system that can enable these speed changes. 
     Rotorcraft propulsion is a critical element of the overall rotorcraft. Unlike fixed wing aircraft, the rotor propulsion system provides lift and control as well as forward thrust. As a result, the rotorcraft engine-gearbox system must be highly reliable and efficient. In addition, the gearbox system must be kept at minimum weight. Presently, the propulsion system accounts for up to 25% of empty vehicle weight. The drive system accounts for up to 72% of the total propulsion system weight. Future rotorcraft trends call for more versatile, efficient, and powerful aircraft, all of which challenge state-of-the-art propulsion system technologies. Variable speed rotors have been identified as having a large impact on many critical rotorcraft issues. 
     Currently, rotor speed can only be varied a small percentage by adjusting the speed of the engine. The variation in rotor speed is generally limited by engine efficiency and stall margin, permitting speed changes limited to approximately 15% when used in current tilt-rotor applications. 
     There is a need for a transmission with a high-range ratio (1:1) for hover mode operation and low-range reduction ratio, such as for example 50% (2:1), through a speed change mechanism, for cruise mode operation. A transmission of this type could be incorporated as an element within the overall propulsion system resulting in overall ratios of 50:1 to 100:1 in the aircraft. 
     It is commonly recognized that variable speed propulsion is required for the design of future advanced rotorcraft. Reductions in rotor speed are required to limit the advancing rotor tip speed and reduce rotor noise. 
     RELATED PATENTS 
     The following patents are incorporated by reference in their entirety herein. 
     U.S. Pat. No. 7,044,877 to Ai discloses for example a two speed transmission having an input shaft and an output shaft. The two-speed transmission is capable of changing the rotating speed of the output shaft from a first speed ratio to a second speed ratio. The shift between the first rotating speed ratio and the second rotating speed ratio is smoothly accomplished by the combination of two sets of planetary gear clusters and two electric motors. The electric motors being are to smooth the mechanical shift between the first speed ratio and the second speed ratio. However, Ai does not disclose a transmission with a high-range ratio (1:1) and low-range reduction ratio, such as for example 50% (2:1), which changes from one to the other through a speed change mechanism including gears, a clutch and a sprag. 
     U.S. Pat. No. 6,641,365 to Karem disclosed for example “A variable speed helicopter tilt rotor system and method for operating such a system are provided which allow the helicopter rotor to be operated at an optimal angular velocity in revolutions per minute (RPM) minimizing the power required to turn the rotor thereby resulting in helicopter performance efficiency improvements, reduction in noise, and improvements in rotor, helicopter transmission and engine life.” 
     US Patent Application No. 2007/0205321 to Waide discloses for example gearboxes providing first and second power-balanced paths in which a speed changer is configured to operate with only one path. Most preferably, the gearbox includes a friction clutch and a sprag clutch arranged such that, together with a lay-shaft and spur-gear differential, gear shifting can be done while transmitting power. The speed changing gearbox of the &#39;321 application has first and second independently and concurrently operational drive paths for transmission of torque. However, Waide does not disclose a transmission with a high-range ratio (1:1) and low-range reduction ratio, such as for example 50% (2:1), which changes from one to the other and directs the torque through an output shaft which is in the same drive path as the input shaft. 
     SUMMARY OF THE INVENTION 
     With the present invention, a transmission, preferably for a rotorcraft is provided where the rotation of the rotor blades can be at 50% or less while maintaining engine speed at the optimal efficiency/performance speed. A portion of the overall 50% reduction can come from extending the engine speed operability range beyond present 15% decrease with the balance provided by the transmission of the present invention. At the present time, the reduction in rotor speed of about 15% is presently accomplished by changing the engine speed. However, with the transmission of the present invention, the entire 50% decrease in rotational speed can be realized without requiring any additional reduction in the speed of the engine. Future overall propulsion system (engine, driveline, and rotor) studies will determine what portion the transmission device should provide for overall optimal performance. This invention uniquely provides both a high-speed 1:1 range and a low-speed 2:1 range (50% speed reduction) with minimal robust parts. The low range ratio being dependent upon the gearing contained within can be varied, as required, to meet specific requirements. 
     According to the present invention, there is disclosed a transmission having a gear arrangement for transmitting torque from an input shaft to an output shaft. The input shaft rotates about a first rotational axis and has a first gear coupled thereto. An elongated, hollow, cylindrical shaft rotates about a second rotational axis that is offset from the first rotational axis. The hollow cylindrical shaft has a second gear at one end there of which meshes with the first gear and a third ring gear at an opposite end thereof. A fourth gear is mounted to one end of a hollow drive shaft. The fourth gear and the hollow drive shaft rotate about the first rotational axis. The fourth gear is meshed with the third gear of the hollow, cylindrical shaft. The hollow drive shaft has a cylindrical end portion at an opposite end thereof. The output shaft rotates about the first rotational axis and has a flange portion attached thereto. A sprag clutch has an input side mounted to the cylindrical end portion of the hollow drive shaft and an output side mounted to the flange portion of the output shaft. A multi-plate clutch is attached to an end portion of input shaft and to the output shaft. Coupling structure is provided for coupling the input shaft with the output shaft whereby the transmission operates in first and second modes. 
     Further according to the present invention, the first mode of operation results in a rotating speed ratio R 1  of 1 to 1 between the input shaft and the output shaft and the second mode of operation results in the rotating speed reduction ratio range of 4.00&gt;R 2 &gt;1.50 between the input shaft and the output shaft. Preferably, the second mode of operation results in the rotating speed reduction ratio range of 2 to 1 between the input shaft and the output shaft. 
     Still further according to the present invention, coupling structure for coupling the input shaft with the output shaft can cause the rotational speed of the output shaft to be the same as the rotational speed of the input shaft is the multi-plate clutch. 
     Yet further according to the present invention, the clutch is a multi-plate clutch having first spaced clutch plates driven by an end portion of the input shaft and second spaced clutch plates which drive the output shaft and interspersed between the first spaced clutch plates. 
     Moreover, according to the present invention, a clutch actuator means engages or disengages the first and second interspersed clutch plates whereby if the multi-plate clutch is engaged the rotational speed of the output shaft is at a first speed which is the same as that of the input shaft and if the clutch is disengaged the rotational speed of the output shaft is at a speed that is different from that of the input shaft. 
     Also, according to the present invention, the multi-plate clutch causes the rotational speed of the output shaft to be the same as the rotational speed of the input shaft whereby the transmission operates in first mode (high speed range, 1:1 ratio). 
     Also, according to the present invention, the sprag clutch causes the rotational speed of the output shaft to be less than the rotational speed of the input shaft whereby the transmission operates in second mode (low speed range, 2:1 ratio). 
     According to the present invention, the first gear has external teeth adapted to mesh with the internal teeth of the second gear and third gear has external teeth adapted to mesh with the internal teeth of fourth gear. 
     Further according to the present invention, the relationship between the output rotational speed and the input rotational speed for the second mode of operation is given by the equation 
               Output   ⁢           ⁢   speed     =     Input   ⁢           ⁢   speed   ×     (       N   14       N   30       )     ×     (       N   34       N   18       )             
where N 14  is equal to the number of teeth on first gear, N 30  is equal to the number of teeth on the second gear, N 34  is equal to the number of teeth on the third gear, and N 18  is equal to the number of teeth on the fourth gear.
 
     Still further according to the present invention, the input shaft is driven by a device from which it receives rotational power such as an engine or an intermediate drive coupling if the present invention is added to an existing design engine-transmission driveline and used as a supplemental inline speed change device. 
     Yet further according to the present invention, the first input shaft of the transmission is connected to the output shaft of a second gear arrangement for transmitting torque from a second input shaft to the first input shaft. The second gear arrangement comprises a second input shaft rotating about the first rotational axis and having a first gear coupled thereto. An elongated, hollow, cylindrical shaft rotating about the second rotational axis is offset from the first rotational axis. The hollow cylindrical shaft having a second gear at one end thereof which engages the first gear and a third gear at an opposite end thereof. A fourth gear is supported by a bearing at the aft end of the input shaft of the transmission. 
     Still further according to the present invention, the relationship between the output speed and input for low speed operation of the second gear arrangement is given by the equation:
 
Output Speed=Input Speed×( N   414   /N   430 )×( N   434   /N   418 )
         Where N 14  is equal to the number of teeth on first gear ( 414 ),   N 30  is equal to the number of teeth on the second gear ( 430 ),   N 34  is equal to the number on the third gear ( 434 ),   and N 18  is equal to the number of teeth on the fourth gear ( 418 ).       

     Yet further according to the present invention, the transmission is a rotorcraft transmission of a light weight configuration with reduced parts. 
     According to the present invention, there is disclosed a method of transmitting torque from an input shaft to an output shaft of a transmission. The method includes the steps of rotating the input shaft having a first gear coupled thereto about a first rotational axis; rotating an elongated, hollow, cylindrical shaft about a second rotational axis that is offset from the first rotational axis, the hollow cylindrical shaft having a second gear at one end thereof which engages the first gear and a third gear at an opposite end thereof; rotating a fourth gear mounted to one end of a hollow drive shaft with a cylindrical end portion at an opposite end thereof about the first rotational axis whereby the fourth gear engages the third gear of the hollow, cylindrical shaft; rotating the output shaft with a flange portion attached thereto about the first rotational axis; mounting an input side of a sprag clutch to the cylindrical end portion of the hollow drive shaft and an output side of the sprag clutch to the flange portion of the output shaft; and coupling the input shaft with the output shaft whereby the transmission ( 10 ) operates in first or second modes. 
     Further according to the present invention, there is disclosed the steps of operating in the first mode of operation resulting in a rotating speed ratio R 1  of 1 to 1 between the input shaft and the output shaft; and operating in the second mode of operation resulting in the rotating speed ratio R 2  of and the second mode of operation results in the rotating speed reduction ratio range of 4.00&gt;R 2 &gt;1.50 between the input shaft and the output shaft. 
     Still further according to the present invention, means are provided for attaching first spaced clutch plates of a multi-plate clutch to an end portion of input shaft and attaching second spaced clutch plates to the output shaft whereby the second clutch plates are interspersed between the first spaced clutch plates; and engaging the first and second interspersed clutch plates whereby the rotational speed of the output shaft is the same as that of the input shaft or disengaging the first and second interspersed clutch plates whereby the rotational speed of the output shaft is less than that of the input shaft. 
     Also according to the present invention, there is disclosed a method of transferring torque from an input shaft to an output shaft of a transmission including the steps of operating in the first mode of operation resulting in a rotating speed ratio R 1  of 1 to 1 between the input shaft and the output shaft; and operating in the second mode of operation resulting in the rotating speed ratio R 2  of 2 to 1 between the input shaft and the output shaft. 
     Further according to the present invention, there is disclosed a method of transferring torque from an input shaft to an output shaft of a transmission including the step of engaging or disengaging the multi-plate clutch whereby when the multi-plate clutch is engaged the rotational speed of the output shaft is at a first speed which is the same as that of the input shaft and when the multi-plate clutch is disengaged the rotational speed of the output shaft is at a speed that is less than that of the input shaft. 
     Also according to the present invention, there is disclosed the steps of engaging or disengaging the clutch whereby when the clutch is engaged the rotational speed of the output shaft is at a first speed which is the same as that of the input shaft and when the clutch is disengaged the rotational speed of the output shaft is at a speed that is less than that of the input shaft. 
     Still further according to the present invention, there is disclosed the steps of coupling the input shaft with the output shaft to cause the transmission to operate in the first mode with the rotational speed of the output shaft the same as the rotational speed of the input shaft is with the multi-plate clutch. 
     Still further according to the present invention, there is disclosed the steps of coupling the input shaft with the output shaft to cause the transmission to operate in the second mode with the rotational speed of the output shaft less than the rotational speed of the input shaft is with the sprag clutch where the output speed is governed by the overall ratio of the gear set comprised of the first, second, third and fourth gears. 
     Yet further according to the present invention, there is disclosed the step of connecting the input shaft to a device that transmits rotational power. 
     Further according to the present invention, there is disclosed the step of serially connecting a plurality of gear arrangements and determining the overall output ratio of the two serially connected gear arrangements from the product of the two in-series ratios, R 1 , R 2 . 
    
    
     
       BRIEF DESCRIPTION OF THE DRAWINGS 
       Reference will be made in detail to embodiments of the disclosure, examples of which may be illustrated in the accompanying drawing figures (FIGURES). The figures are intended to be illustrative, not limiting. Although the invention is generally described in the context of these embodiments, it should be understood that it is not intended to limit the invention to these particular embodiments. 
       Certain elements in selected ones of the figures may be illustrated not-to-scale, for illustrative clarity. The cross-sectional views, if any, presented herein may be in the form of “slices”, or “near-sighted” cross-sectional views, omitting certain background lines which would otherwise be visible in a true cross-sectional view, for illustrative clarity. In some cases, hidden lines may be drawn as dashed lines (this is conventional), but in other cases they may be drawn as solid lines. 
       If shading or cross-hatching is used, it is not intended to be of use in distinguishing one element from another (such as a cross-hatched element from a neighboring un-shaded element). It should be understood that it is not intended to limit the disclosure due to shading or cross-hatching in the drawing figures. 
         FIG. 1  is an oblique cross sectional view of a two-speed, mechanical-power-conveying transmission, according to the present invention. 
         FIG. 2A  is an orthogonal cross sectional view of the two-speed, mechanical-power-conveying transmission, according to the present invention. 
         FIG. 2B  is a schematic axial view of the gear relationships in the two-speed, mechanical-power-conveying transmission, in according to the present invention. 
         FIG. 3A  is an orthogonal cross sectional view of the present invention showing the path of power flow during high-speed output operation of the two-speed, mechanical-power-conveying transmission, according to the present invention. 
         FIG. 3B  is an orthogonal cross sectional view of the present invention showing the path of power flow during low-speed output operation of the two-speed, mechanical-power-conveying transmission, according to the present invention. 
         FIG. 4  is an orthogonal cross sectional view of multiple stages of the two-speed, mechanical-power-conveying transmission in series, according to the present invention. This configuration consists of a fixed-ratio first stage and a variable-ratio second stage with only the second stage clutch controlled. 
         FIG. 5  is an orthogonal cross sectional view of multiple stages of an alternate two-speed, mechanical-power-conveying transmission in series, according to the present invention. This configuration consists of two variable-ratio stages both simultaneously clutch controlled. 
         FIG. 6  is an orthogonal cross sectional view of the above ( FIG. 5 ) in a configuration in which the power input is transferred directly into the input gear in lieu of the input shaft end providing for up to three outputs at different speeds, both of fixed ratio and variable ratio relative to the motor input speed. 
     
    
    
     DETAILED DESCRIPTION OF THE INVENTION 
     Throughout the descriptions set forth herein, various features of the invention may be described in the context of a single embodiment. The features, however, may also be provided separately or in any suitable combination. Conversely, although the invention may be described herein in the context of separate embodiments for clarity, the invention may also be implemented in a single embodiment. Furthermore, it should be understood that the invention can be carried out or practiced in various ways, and that the invention can be implemented in embodiments other than the exemplary ones described hereinbelow. The descriptions, examples, methods and materials presented in the description, as well as the claims, should not be construed as limiting, but rather as illustrative. 
     If any dimensions are set forth herein, they should be construed in the context of providing some scale to the relationship between the elements. For example, a given element may have an equal, lesser or greater dimension (such as diameter) than another element. Any dimensions that are important or critical will generally be identified as such. The term “at least” includes equal to or greater than. The term “up to” includes less than. Also, an open-ended range or ratio as “at least 2:1” should be interpreted to include sub-ranges such as at least 2:1, at least 5:1, and at least 10:1. 
     The present two-speed transmission invention  10  is referred to herein as a “transmission,” “two-speed transmission,” “compound gear transmission,” or variations thereof, or as the inventors&#39; preferred usage: “offset compound gear drive,” or OCG. 
     Referring to  FIG. 1  there is shown in cross sectional view the novel two-speed, mechanical-power-conveying transmission  10  comprising an input shaft  12  having a first gear  14  attached thereto, an output shaft  26 , an elongated, hollow, cylindrical shaft  16  having a second gear  30  possessing internal teeth at one end  16   a  of the cylindrical shaft, and a third gear  34  possessing external teeth at the other end  16   b  of the cylindrical hollow shaft, a fourth gear  18  integral with or attached to wheel  65  and supported on bearing  62  maintaining concentricity with gear  14  by means of the aft end of shaft  12 , a multi-plate clutch  22 , and a sprag clutch  28 . The input shaft ( 12 ) is rotationally driven by a device ( 13 ), such as an engine, from which it receives rotational power. The elongated cylindrical shaft  16  has a rotational axis  19  that is offset from a shared main rotational axis  21  of the input shaft  12  and the output shaft  26 . Input shaft  12  is supported on bearings  60  and  62 . Output shaft  26  is supported on bearings  64  and  66 . Bearings  60  and  66  serve as drive system main bearings. Bearings  62  and  64  serve as intermediate bearings. Bearing  62  maintains concentricity and permits relative differential rotational speeds between gear  18  and shaft  12 . Bearing  64  serves as a pilot bearing between input shaft  12  and output shaft  26  to maintain concentricity and permit relative differential rotational motion between input shaft  12  and output shaft  26  (i.e., differential rotational speeds). Bearings  60 ,  62 ,  64 , and  66  share a common central axis  21 . Shaft  16  is supported on bearings  48  which are concentric with the axis  19  and are offset from the central axis  21  of bearings  60 ,  62 ,  64 , and  66 . The offset between the axis  19  and  21  is a direct function of the gear ratios. Bearings  60 ,  62 ,  64 , and  66  are represented as rolling element type bearings but may also be of the fluid film type or magnetic type as warranted by overall transmission speed and power requirements, and bearings  48  which are represented as fluid film type journals may also be of the rolling element type. 
     The first gear  14  having external teeth, is affixed to the input shaft  12 , and meshes with the second gear  30  having internal teeth, on the forward end  16   a  of the elongated hollow shaft  16 . At the aft end  16   b  of the elongated hollow shaft  16  is a third gear  34 , having external teeth, and which meshes with a fourth gear  18 , having internal teeth. The fourth gear  18  is attached to a hollow drive shaft  20  at a forward end  20   a , by means such as bolts  29  (see  FIG. 2A ). The opposite or distal end  20   b  of hollow drive shaft  20  has a cylindrical end portion  20   c  onto which is mounted the input side  28   a  of a sprag clutch  28 . 
     The output shaft  26  has an integral flange portion  43  that is located between opposite ends  26   a  and  26   b  of the output shaft  26 . Flange portion  43  has an upstanding rim  43   a  with an inner surface  43   b  that receives the output side  28   b  of the sprag clutch  28 . The input shaft  12  has an end portion  12   b  that is large in diameter as compared to the remainder of the input shaft  12 . 
     A hollow cylinder  24 , sized to accommodate the multi-plate clutch assembly  22 , is attached or contiguous with the input shaft distal end portion  12   b  of input shaft  12 . The clutch assembly  22  has alternating stacked clutch plates  22   a ,  22   b ,  22   c ,  22   d  ( 22   a - 22   d ) (see  FIG. 2 ) that are driven by, and rotate at the same speed as hollow cylinder  24  by means of a spline or tooth engagement at the outer perimeter. The spaced stacked plates  22   a - 22   d  engage by means of friction a set of interspersed stacked clutch plates  23   a ,  23   b ,  23   c ,  23   d  ( 23   a - 23   d ) which drive and rotate at the same speed as the output shaft  26  by means of spline or tooth engagement at the inner perimeter. An annularly arranged clutch actuator  51 , which is mounted to and rotates with the end portion  12   b  of input shaft  12 , compresses or releases the clutch  22  to cause it to engage or disengage during operation, as described herein below. The configuration of the clutch actuator  51  is a mechanical spring arrangement (e.g., helical coil, Belleville, diaphragm spring) activated and hydraulically released (e.g., by an annular piston). A mechanical fail safe feature is incorporated in the clutch release (disengagement) mechanism so that the clutch will be engaged if there is a failure of the clutch release mechanism.  FIG. 2B  provides an axial schematic view of the rotating components of the present transmission invention  10 . The elements shown are the input shaft  12  (which has the output shaft  26  behind it and out of view), having the first gear  14  attached thereto, a ball or roller type bearing  17 , the second gear  30  and the third gear  34  that are part of, and integral with, the offset hollow cylindrical shaft  16 , and the fourth gear  18 . The solid line  27  defines the foreshortened, end view of the cylindrical surface plane of bearings  48 , which provide support to offset hollow shaft  16 . The offset aspect of the hollow driveshaft  16  is evident in the location of its axis of rotation  19  in relation to the axis of rotation  21  that is shared by the input shaft  12  and the output shaft  26 . Axis of rotation  21  is the central, or primary, machine axis on which the drive system input and output are centered, whereas axis of rotation  19  is a secondary axis of rotation on which some of the internal components between the input and output rotate, primarily hollow shaft  16 , second gear  30 , and third gear  34 . The dashed oval  37   a  encompasses a first mesh plane  37  where the first gear  14  meshes with the second gear  30 , and the second dotted oval  41   a  encompasses the second mesh plane  41  where the third gear  34  meshes with the fourth gear  18 .  FIG. 2B  in an idealized view combining mesh plane  37  and mesh plane  41  into a single plane for presentation of the OCG Offset Compound Gear concept basis, whereas in the present invention the two mesh planes are separated axially. 
     A first bearing set  60  (see  FIGS. 1 AND 2A ) supports the input end  12   a  of the input shaft  12 . A first single bearing  62  supports the fourth gear  18  in relation to the input shaft  12 . A second single bearing  64  supports the output shaft  26  in relation to the input shaft  12 . A second bearing set  66  supports the output shaft  26 . The hollow, cylindrical offset shaft  16  is carried by bearings  48 , which are of the fluid film journal/thrust type or rolling element bearing type based upon specific transmission requirements. 
     Operational Dynamics 
     During operation, if the multi-plate clutch  22  is engaged, then the rotational speed of the output shaft  26  is the same as that of the input shaft  12  and the power flows directly from the input shaft  12  to the output shaft  26  through the multi-plate clutch  22  by means of torque transmitted via friction created by the clamping force provided by releasing clutch actuator  51 . If the clutch  22  is disengaged, then the rotational speed of the output shaft  26  is less than that of the input shaft  12  and the power flows from the fourth gear  18  to the flange portion  43  of the output shaft  26  by way of the sprag clutch  28 . The ratio of the input rotational speed and the output rotational speed when the clutch  22  is disengaged is on the order of 2:1 as described or some other ratio as required. 
     The input/output speed ratio is a function of the effective respective diameters of the first and second meshing gears  14  and  30 , respectively, and the respective diameters of the third and fourth meshing gears  34  and  18 , respectively, as should be readily evident to those who are skilled in the art of transmission of rotary mechanical power. The input/output ratio is discussed in more detail hereinbelow. 
     The two-speed operation of the present transmission invention  10  becomes more evident upon contemplation of cross sectional views of  FIGS. 3A and 3B .  FIG. 3A  illustrates high-speed operation of the present transmission invention  10 , which takes place when the multi-plate clutch assembly  22  is engaged. The direction of flow of rotary mechanical power is shown by means of the line  77  with arrowheads  77   a . The direction of flow of mechanical rotary power is from input shaft  12 , hollow cylinder  24 , to clutch assembly  22  and through output shaft  26 , such that the output speed is the same as the input speed (the output ratio is 1:1.) 
       FIG. 3B  illustrates low-speed operation of the present transmission invention  10 , which takes place when the multi-plate clutch assembly  22  is disengaged. Power enters at the input shaft  12  and is transferred by way of the first gear  14  to the second gear  30 , which is contiguous with the hollow driveshaft  16 . The hollow driveshaft  16  conveys rotary power to the contiguous third gear  34 , which transmits it to the fourth gear  18 , which conveys it onward to the hollow drive shaft  20  that is affixed, such as by means of bolts  29 , to a wheel portion  65  of the fourth gear  18 . The hollow driveshaft  20  conveys power to the sprag clutch  28 , which transmits it onward to the flange portion  43  of the output shaft  26 , such that the output speed is less than the input speed, such as for example the output ratio is 2:1. 
     Input/Output Speed Ratios 
     The cylindrical offset shaft  16  comprises second and third gears  30  and  34 , respectively, disposed respectively at opposing ends  16   a  and  16   b  of the offset hollow shaft assembly. The second gear  30  has internal gear teeth  30 ′ and the third gear  34  has external gear teeth  34 ′. The second gear  30  of the cylindrical offset shaft  16  receives mechanical rotary power from the first gear  14  at the first gear mesh  52  ( FIG. 2A ) and then conveys the rotary mechanical power by means of the third gear  34  that meshes with the fourth gear  18  at the second gear mesh  54 . 
     The internal gear teeth  30 ′ of second gear  30  of the cylindrical offset shaft  16  receive rotary force from the external gear teeth  14 ′ of the first gear  14 ; the external teeth  34 ′ of the third gear  34  conveys rotary force to the internal gear teeth  18 ′ of the fourth gear  18  which conveys rotary mechanical power to the hollow driveshaft  20 . Input/output speed ratios are determined by the respective numbers of gear teeth  30 ′,  34 ′,  14 ′,  18 ′ of the two meshing gear sets, first and second gears  14 ,  30 , respectively and third and fourth gears  34 ,  18 , respectively. 
     The respective gear teeth  30 ′,  34 ′,  14 ′,  18 ′ of first and second gears  14 ,  30 , respectively and third and fourth gears  34 ,  18 , respectively, can be of the straight cut spur varieties or of the helically cut or other gear teeth types such as herringbone as deemed necessary for required power rating operational reliability and quiet operation. 
     Note that all rotating parts described hereinabove rotate in the same direction. Reductions in rotary speed take place at two locations: (1) at the first gear mesh  52  between the first gear  14  and the second gear  30  of the cylindrical offset, double-gear assembly  16  and, (2) at the second gear mesh  54  between the third gear  34  of the second gear  18 . 
     The relationship between the output speed and input for the low speed operation is given by 
               Output   ⁢           ⁢   speed     =     Input   ⁢           ⁢   speed   ×     (       N   14       N   30       )     ×     (       N   34       N   18       )             
where N 14  is equal to the number of teeth on first gear  14 , N 30  is equal to the number of teeth on second gear  30 , N 34  is equal to the number of teeth on third gear  34 , and N 18  is equal to the number of teeth on fourth gear  18 .
 
     The remainder of this section is a discussion on the ratio range potential of the OCG. The term “R” means the same as “ratio.” 
     The ratio-range potential for the speed reduction between the input and the output shafts  12  and  26  of the OCG transmission  10  in a single-stage configuration is 4.00&gt;R&gt;1.50 (speed reduction output) and, conversely, it is 0.25&lt;R&lt;0.75 for a back driven, or reverse (speed increasing output) configuration. Preferably, however, the speed reduction ratio from the input to the output is 2:1 or R=2.0. 
     OCGS in Series: Arrangement One (Multi-Stage Fixed/Variable Ratio) 
     Referring now to  FIG. 4 , there is shown, in cross-sectional schematic view, two OCGs  310 ,  410  coupled in such a way that the output of a first OCG  310  is directed into a second OCG  410  so as to provide a series arrangement  300  wherein the overall ratio of input/output speed reduction (or multiplication) can be greater than that of a single OCG. It is possible, as those skilled in the art would clearly appreciate, that an unlimited number of OCGs could be so serially arranged, though practical considerations would necessarily place limits. 
     The series arrangement  300 , portrayed in  FIG. 4 , includes the first OCG  310  which is comprised of the gear portion only of the transmission  10  described hereinabove. The first OCG  310  has an input shaft  312  and three moving parts with gears such that the input shaft drives a fifth gear  314 , a hollow driveshaft  316 , and an eight gear  318 , which correspond respectively to the first gear  14 , the hollow driveshaft  16  and the fourth gear  18  in the above described OCG transmission  10 . The operational dynamics of the OCG gear train  310  need not be described again, as it is the same as that given hereinabove in relation to the basic OCG transmission  10 . 
     First OCG  310 , as shown in  FIG. 4 , has an output shaft formed of a flange  313  which is secured to the eighth gear  318  by means such as screws  329 . The output shaft  413  is shown as being contiguous with, and/or is one in the same as, the input shaft  412  of the second OCG portion  410 . The other parts of the second OCG portion  410 , and their operational dynamics, are as described hereinabove in reference to the OCG transmission  10 . It should be noted that the method of bolting the output shaft  313  to eighth gear  318  is only one of many such coupling methods that could be used to greater or lesser advantage in the series coupling of the present OCG series arrangement  300 . Splined connections could be used, or other types of bolted couplings, including flexible or universal joints could also be used, as called for by competent engineering judgment. 
     The second OCG portion  410  consists of the input shaft  412 , a fifth gear  414  (compare first gear  14 ), a hollow driveshaft  416 , an eight gear  418  (compare fourth gear  14 ), a hollow driveshaft  420  housing a clutch  422 , a sprag clutch  428 , and an output shaft  426 , each of which, with the exception of the input shaft having the flange  413  has corresponding parts as described hereinabove in reference to the OCG transmission  10 . That is to say, the second OCG transmission portion  410  displayed in  FIG. 4  is of the same physical and operational sort that is described hereinabove as the transmission invention  10 . 
     In operation, the overall series arrangement  300  provides an overall rotational speed reduction between the input shaft  312  and the output shaft  426  that is the multiplicative product of the speed reduction ratio of the first OCG portion  310  and the speed reduction ratio of the second OCG portion  410 . Thus the input/output speed reduction ratio exceeds that of a single OCG transmission  10 . Note also that said speed reduction property could, upon reverse driving, provide a speed multiplication, as should be obvious to those who are skilled in the art. 
     In the embodiment shown in  FIG. 4 , ratios above 4.00 can be configured using multiple stages of the transmission  10  in series. As an example, the OCG ratio range potential for the OCG Drive  10  in a two-stage configuration can provide R=16.00 by employing two R=4.00 (as the single stage limit) in series (i.e., 16=4×4). Ratios between 4 and 16 can be created using a combination of twin ratios or dual ratios in series. A series arranged dual-ratio, two-stage configuration would employ configurations with two different OCG ratios to provide the desired overall output ratio. The overall ratio is defined as the product of the two in-series ratios, R 1 , R 2  (i.e., R out =R 1 ×R 2 ). The subject of two-stage configurations does not imply the use of two duplicate configurations, each with a clutch and sprag. More properly, it means that the OCG gearing would be employed in a two-stage configuration while maintaining the single clutch and sprag. 
     For a speed reduction of the sort the OCG was designed for, the upper and lower ratio limits are restricted by the geometrically possible input gear size. Above R=4.00, the input gear becomes impractically small because the bearing size and shaft become impossibly small. Below R=1.50, the input gear becomes too large creating gear tooth interference with the internal teeth of the second mesh. 
     OCGS in Series: Arrangement Two (Multi-Stage Variable/Variable Ratio) 
     Referring now to  FIG. 5 , there is shown, in cross-sectional schematic view, two OCGs  310 A and  410 A, coupled in such a way that the output of a first OCG  310 A is directed into a second OCG  410 A so as to provide an alternate series arrangement  300 A, wherein the overall ratio of input/output speed reduction (or multiplication) can be also be greater than that of a single OCG. In contrast to the series arrangement described above, and shown in  FIG. 4 , the coupling of the two OCGs,  310 A and  410 A, as shown in  FIG. 5 , is configured to change the power flow through both OCGs  310 A and  410 A simultaneously, permitting a larger overall output speed ratio change though the OCG device than that of  FIG. 4 . In  FIG. 4 , the series arrangement only permits ratio change in the second OCG  410 . In  FIG. 4 , OCG  310  is used as a fixed ratio device. In  FIG. 5 , OCG  310 A and OCG  410 A are mechanically connected in a manner which permits both OCG to simultaneously shift between output ratio R=1:1 direct drive (clutch engaged), or the combined ratio of OCG  310 A and OCG  410 A (clutch disengaged). The selection of either the configuration described in  FIG. 4 , or that described in  FIG. 5 , is dependent upon design requirements of the intended end use application. 
     The series arrangement  300 A portrayed in  FIG. 5  includes the first OCG  310 A which is comprised of the gear portion only of the transmission  10  described hereinabove. The first OCG  310 A has an input shaft  312 A and three moving parts with gears such that the input shaft drives a gear  314 , a hollow driveshaft  316 , and gear  318 , which correspond respectively to the gear  14 , the hollow driveshaft  16  and gear  18  in the above described OCG transmission  10 . The operational dynamics of the OCG gear train  310 A need not be described again, as it is the same as that given hereinabove in relation to the basic OCG transmission  10 . 
     Differences relative to OCG  300  shown in  FIG. 4 , reflected in OCG  300 A shown in  FIG. 5 , provide a different functionality, and consists of several changes. First, combining shaft  312  and shaft  412 (ref:  FIG. 4 ) into an integral shaft  312 A, which is now common to both OCG  310 A and  410 A. Secondly, flange  313  (ref:  FIG. 4 ) is eliminated and is replaced with drive hub  414 A (forward extension of gear  414 ). Thirdly, gear  414  is mechanically decoupled at  414 B from shaft  312 A, and is now supported by bearings located between hub  414 A at gear  414  and shaft  312 A, permitting relative rotational motion between gear  414  and shaft  312 A, with no transfer of power at  414 B. The effect of the above three reconfiguration changes redirects power from OCG  310 A (gear  318 ) directly to gear  414  via drive hub  414 A. Drive hub  414 A is depicted as integral to gear  414 . In certain instances, it may be advantageous for hub  414 A and gear  414  to be separate parts mechanically connected to provide power transmission capability in lieu of being a single integral part as shown in  FIG. 5 . Whether hub  414 A and gear  414  are integral in configuration, or two separate parts mechanically connected, their function in  FIG. 5  is to transfer power from ring gear  318  directly to input gear  414  of second stage OCG  410 A, which are mechanically connected via bolts  329 . 
     Relative to  FIG. 5 , the second OCG portion  410 A consists of the alternate configuration input shaft  312 A (common to, and shared by both OCG  310 A and OCG  410 A), a gear  414  (compare gear  14 , except changes described above regarding decoupling of gear  414  from shaft  312 A at  414 B), a hollow driveshaft  416 , a gear  418  (compare gear  14 ), a hollow driveshaft  420  housing a clutch  422 , a sprag clutch  428 , and an output shaft  426 , each of which, with the exception of reconfigured input gear  414  and shaft  312 A, has corresponding parts as described hereinabove in reference to the OCG transmission  10 . That is to say, the second OCG transmission portion  410 A displayed in  FIG. 5  is of the same physical and operational sort that is described hereinabove as the transmission invention  10 , except with power entry directly to gear  414 / 414 A in lieu of input shaft  12 . 
     In operation, the overall alternate series arrangement  300 A provides an overall rotational speed reduction between the input shaft  312 A and the output shaft  426  that is the multiplicative product of the speed reduction ratio of the first OCG portion  310 A and the speed reduction ratio of the second OCG portion  410 A. Thus the input/output speed reduction ratio exceeds that of a single OCG transmission  10 . In addition, this alternate arrangement permits the simultaneous shifting of both OCG  310 A and OCG  410 A resulting in an overall ratio change between 1:1 (clutch engaged) and the multiplicative product of the speed reduction ratio of the first OCG portion  310 A and the speed reduction ratio of the second OCG portion  410 A (clutch disengaged). This functionality is in contrast with that of OCG  300  series depicted in  FIG. 4  where only OCG  410  is clutch controlled and OCG  310  is a fixed-ratio device. 
     In the embodiment shown in  FIG. 5 , ratios above 4.00 can be configured using multiple stages of the transmission  10  in series. As an example, the OCG ratio range potential for the OCG Drive  10  in a two-stage configuration can provide R=16.00 by employing two R=4.00 (as the single stage limit) in series (i.e., 16=4×4). Ratios between 4 and 16 can be created using a combination of twin ratios or dual ratios in series. A series arranged dual-ratio, two-stage configuration would employ configurations with two different OCG ratios to provide the desired overall output ratio. The overall ratio is defined as the product of the two in-series ratios, R 1 , R 2  (i.e., R out =R 1 ×R 2 ). The subject of two-stage configurations does not imply the use of two duplicate configurations, each with a clutch and sprag. More properly, it means that the OCG gearing would be employed in a two-stage configuration while maintaining the single clutch and sprag. The embodiment shown in  FIG. 5 , provides the output ratios 1:1 (direct drive), or R out , where R out =R 1 ×R 2 , whereas in the embodiment shown in  FIG. 4 , the output ratios are R 1 , where R 1  is that of the first-stage OCG, or R out , where R out =R 1 ×R 2 . It should be obvious, that for an equally geared pair of OCG in series that the  300 A configuration as shown in  FIG. 5  provides a greater overall ratio change than OCG  300  in  FIG. 4 . 
     Multiple Output OCG Application 
     Referring now to  FIG. 6 , there is shown, in cross-sectional schematic view, a configuration consisting of an OCG  300 A ( FIG. 5 ), or, optional series drive OCG  300  ( FIG. 4 ), or, optional basic OCG  10  ( FIGS. 1 and 2 ), coupled to a drive motor  500  such that the configuration provides for a possibility of three output speed capability. 
     It should be obvious that the introduction of gear  314 A, which is mechanical coupled to gear  314 , redirects the power entry in the OCG from shaft  312 A to gear  314 . 
     The primary output is located at output shaft  426 , with optional outputs at shaft  504 , and/or at output shaft  312 A, each at a unique speed(s) with respect to drive motor  500 . For the depicted figure, the speeds for the output shafts are as follows. Output shaft  426  rotational speeds are direct drive ratio 1:1, or the speed determined by OCG overall ratio (R1×R2), dependent upon clutch mode. Output shaft  312 A speed is the product ratio of gear  502  and gear  314 A, at a fixed speed relative to the motor  500  speed. Output shaft  504  is direct drive and equal to the speed of motor  500 . Motor  500  may be a fixed or variable speed device as required in the end application. The orientation of motor  500  and/or rotational axis is not limited to the example configuration, but may be placed to the left of gear  502 / 314 A mesh if advantageous, or even perpendicular to the central axis of the OCG using appropriate gear geometry as the end application design requirements dictate. 
     The fixed-ratio speeds available at shaft  312 A and shaft  426  are a function of the drive ratio determined by gear  502  and gear  314 A and the former combined with the overall reduction ratio of OCG  300 A,  300 , or  10 . These gears ( 502  and  314 A) may be speed-increasing as shown by gear  502  and gear  314 A, or speed-decreasing optional gear set comprised of gear  502 A and gear  314 B. The configuration depicted in  FIG. 6  provides a wide range of flexibility in the number of possible output speeds, and their respective speed ratios relative to the speed of motor  500 . Outside the intended application for the OCG drive described in the background, an application for this configuration, utilizing a gear set such as example gear set  502 A/ 314 B, may be a high-reliability long-life low speed/power gear drive system for extraterrestrial applications requiring a clutch-controlled two speed drive and one/multiple power take-offs which may be independently clutch controlled (coupled/uncoupled) and driven by a common motor. 
     If no intermediate output speeds are required for a given application, the user should use configurations Alternate Series OCG Drive  300 A ( FIG. 5 ), or optional series drive OCG  300  ( FIG. 4 ), or basic OCG  10  ( FIGS. 1 and 2 ) coupled to a driving device. 
     Although the invention has been shown and described with respect to a certain preferred embodiment or embodiments, certain equivalent alterations and modifications will occur to others skilled in the art upon the reading and understanding of this specification and the annexed drawings. In particular regard to the various functions performed by the above described components (assemblies, devices, etc.) the terms (including a reference to a “means”) used to describe such components are intended to correspond, unless otherwise indicated, to any component which performs the specified function of the described component (i.e., that is functionally equivalent), even though not structurally equivalent to the disclosed structure which performs the function in the herein illustrated exemplary embodiments of the invention. In addition, while a particular feature of the invention may have been disclosed with respect to only one of several embodiments, such feature may be combined with one or more features of the other embodiments as may be desired and advantageous for any given or particular application.

Technology Classification (CPC): 5