Abstract:
The transmission includes a minimal-orbiter gear complex and a single infinitely-variable rotary control device. The minimal orbiter includes only a control gear and an output gear interconnected by the different gearing portions of at least one cluster gear supported by an orbiting web responsive to an input drive provided by a primary engine. The rotary control device may be any kind of apparatus that is capable of providing resistance torque that can match the torque of the primary engine to slow and stop the control gear of the orbital complex. In a preferred embodiment disclosed, the rotary control device is a hydraulic jack machine having a drive shaft connected to an adjustable swash plate that provides primary control of the flow of fluid through the machine.

Description:
REFERENCE TO RELATED APPLICATIONS 
   The subject matter in this application is related to the subject matter in U.S. patent application Ser. No. 11/153,112, filed Jun. 15, 2005, entitled “ORBITAL TRANSMISSION WITH GEARED OVERDRIVE”, now U.S. Pat. No. 7,475,617, issued Jan. 13, 2009. The aforementioned application is hereby incorporated herein by reference. 
   BACKGROUND OF THE INVENTION 
   1. Field of the Invention 
   The invention pertains to the field of automotive transmissions. More particularly, the invention pertains to an automotive transmission with orbital gearing and a variable sun gear control. 
   2. Description of Related Art 
   Known conventional transmissions use the vehicle&#39;s engine as the primary control for making changes in vehicle speed. 
   The manual transmission uses a clutch to change the gear ratio, with the engine being completely disconnected from the transmission momentarily during each level of gear change, and also with the engine being quickly revved up to fairly high rpm during each level change. 
   The standard automatic transmission uses a torque converter to avoid the complete disconnect of the engine between levels of gear ratio change, but the inefficiency of the torque converter causes considerable slippage between engine and transmission output, particularly during initial start-up and lower speeds when as much as 50% of the engine torque may be lost. This type of transmission blends engine and transmission better than the manual, but the engine still must be revved to higher rpm during each level of multiple gear change. Also, even during engine idle when the vehicle is stationary, the automatic transmission creates a constant loss of efficiency through hydraulic losses occurring in the torque converter. 
   The conventional acceleration rates for engine rpm during the revving for each gearing level in both manual and automatic transmissions is often between 1000-2000 rpm/sec, and this rapid acceleration of the engine&#39;s internal parts (crank shaft, pistons, valve cams) can result in a 20-25% loss in efficiency. 
   There have been many different forms of automatic infinitely-variable transmissions (“IVT”), in which usable torque is supplied to the drive wheels of the automotive vehicle through a continuum of constantly variable speeds. The IVT is distinguished from the continuously-variable transmission (“CVT”), in which vehicle speed is continuously changed throughout several successively-increasing speed and torque output levels. However, until recently, no IVT or CVT has been developed that is capable of successfully handling a full range of torque and engine sizes from a very small vehicle through a large commercial truck. Torvec, Inc., the assignee of the present invention, has recently successfully tested an IVT that can be readily sized to cover this entire range of engine size and torque requirements. Also, this recently tested IVT was specifically designed for propelling not only large SUVs (sport utility vehicles) but also small trucks and school busses. One of the latest designs of the Torvec IVT is disclosed in U.S. Pat. No. 6,748,817 entitled “Transmission with Minimal Orbiter”. 
   Torvec IVT&#39;s have been progressively improved during an extensive period of product testing, and a current design produces continuous changes of torque and speed from start-up through an overdrive ratio without any intermediate discontinuities while using engine acceleration rates of no more than 90-100 rpm/sec. These remarkable results are achieved with an apparatus that is significantly smaller and lighter than presently available conventional automotive transmissions. 
   The earlier Torvec IVT designs combined a variable hydraulic pump and a hydraulic motor with a gear orbiter to form an infinitely variable transmission so that, as the speed of the hydraulic motor increases the rotation of the gear orbiter, the output shaft speed increases and the speed of the vehicle increases. This basic design was recently significantly modified to operate in an unconventional manner. Namely, while the engine input was delivered to an input sun gear of the orbiting gear complex, the changes in output gear ratios were obtained by using the combined operation of the variable hydraulic pump and motor to slow the rotation of the web so that, as the rotation of the web in the direction of the engine was slowed, the transmission produced a continuously decreasing gear reduction, and, when the web was brought to a stop, the transmission provided an overdrive ratio of the engine input. 
   The recent Torvec transmissions just discussed above increase transmission efficiency by using a hydraulic pump and motor combination, rather than the vehicle engine, as a primary control of vehicle speed, thereby avoiding the above-mentioned engine acceleration losses. However, these recent Torvec transmissions still lose efficiency through the split torque path that delivers torque to the hydraulic pump and motor. 
   As just indicated above, when accelerating a vehicle with both manual and automatic transmissions, the revving of the engine, far exceeding 2,000 rpm, during various gear change levels is an inefficient use of horsepower. Even the relatively new Torvec IVT transmissions are burdened with the inefficient use of horsepower that arises from the division of a portion of the vehicle engine output into a split path for driving the transmission&#39;s hydraulics. 
   Special attention is also called to another prior art apparatus, namely, the Torvec long-piston hydraulic machines disclosed in U.S. Pat. No. 6,983,680 and U.S. 2004/0168567, which are hereby incorporated herein by reference. This special prior art is referred to in greater detail in the Summary portion below. 
   The invention disclosed herein is a further improvement of the successfully-tested earlier versions of the Torvec IVT just discussed above, and the hydraulic machine in the disclosed preferred embodiment utilizes a variation of the hydraulic machine disclosed in U.S. Pat. No. 6,983,680 and U.S. 2004/0168567. 
   SUMMARY OF THE INVENTION 
   The inventive transmission is a remarkably small structure that includes a minimal-orbiter gear complex and a single a rotary control device. The minimal orbiter includes only: a control gear and an output gear interconnected by the different gearing portions of at least one cluster gear supported by an orbiting web responsive to an input drive provided by a primary engine. The rotary control device may be any kind of apparatus that is capable of providing infinitely-variable resistance torque sufficient to match the torque of the primary engine to slow and stop the control gear of the orbital complex. In a preferred embodiment, the rotary control device is a hydraulic jack machine having a drive shaft connected to an adjustable swash plate and having input and output ports connected through a very minimal fluid passage that is closed by a controlled pressure valve. 
   While the disclosed preferred embodiments of a transmission of the present invention all utilize a single hydraulic machine as the rotary control device, all of these embodiments omit the split torque path that is conventionally used in hydraulically-controlled transmissions. None of the disclosed embodiments directly split off a portion of the engine torque to a path for operating the hydraulics. Instead, the engine torque is directed only through the mechanical gearing of the transmission. Also, the single hydraulic machine that generates resistance torque during vehicle acceleration places negligible load on the vehicle&#39;s engine during start-up, engine idle, and vehicle cruising. 
   To summarize the preferred operation of a transmission of the present invention: When the vehicle&#39;s engine is initially started, the orbital web of the small gear complex moves with the engine drive. The transmission output gear is connected to the vehicle&#39;s drive shaft, and when the parked vehicle&#39;s wheels are standing still on the terrain, the output gear is held in a stopped condition, and the cluster gears spin around the stopped output gear as the orbital web moves with the engine drive. Under these conditions with the preferred gear ratios indicated herein, the control gear rotates at approximately one-half the engine input speed. The swash plate of the hydraulic machine is set at 0°; the control valve is open; and the hydraulic drive gears and the shaft of the hydraulic machine merely rotate freely with the control gear at one-half engine speed, adding only a minimal frictional load. 
   In this regard, special attention is called to the fact that preferred embodiments of the invention use a variation of the above-mentioned prior art Torvec long-piston hydraulic machine. Commercial-quality prototypes of these Torvec long-piston hydraulic machines have already been successfully built and tested in both large SUVs (sport utility vehicles) and small trucks, and while these remarkable hydraulic machines have not yet enjoyed wide publicity, they are the preferred hydraulic machines for use with the invention described herein. The preference for this design of hydraulic machine cannot be over emphasized, since presently available commercial hydraulic pumps and motors are considered unacceptable for use with the subject invention because: (1) they are much larger and much heavier than the Torvec long-piston machines; (2) they are incapable of providing the high speeds needed for automotive use; (3) they do not have the start-up torque capabilities of the Torvec long-piston machines; (4) their “break-away” torque makes them inappropriate for the invention, requiring tens of pounds of force to begin to turn their drive shafts even when unloaded, whereas the unloaded drive shaft of a Torvec long-piston machine can be rotated by hand or finger grip; and (5) the volumetric efficiency of present commercial hydraulic machines is poor at low swash plate angles, whereas in actual testing, a Torvec machine produced 2000 psi or more at a swash plate angle of 1.5° with an input speed of 1700 rpm, registering a volumetric efficiency of about 95% at this small angle. With these just-listed deficiencies, if such standard hydraulic machines were to be used in the subject transmission, many of the advantages of the invention would be lost, e.g., the invention&#39;s neutral “no-load” condition could not be achieved, the vehicle&#39;s brake would have to be applied to avoid vehicle “creep” when standing in traffic on level ground, etc. 
   While the orbital gearing of a transmission of the present invention is preferably connected at all times to the hydraulic machine, the only notable load provided by the hydraulics occurs when the swash plate is adjusted to change the transmission gear ratios during vehicle operation. This hydraulic load provides a resistance torque that gradually slows the speed of rotation of the control gear through a continuum of decreasing speeds that begin when the swash plate is tipped as little as 1-1.5°. The progressive increase in resistance torque generated by the hydraulic machine causes a proportionally progressive slow-down of the control gear. The slow down of the control gear creates driving torque from the transmission output through an infinitely-varying gear ratio that begins, momentarily, from ∞:1-300:1 when the swash plate is tipped as little as 1-1.5°, and ends when the swash plate is at 25°, bringing the control gear to a stop and producing a transmission output at a predetermined overdrive ratio (e.g., 1:0.7). 
   Special attention is called to the fact that a hydraulic machine is not acting like a conventional hydraulic pump or motor in the present invention. Thus, the increasing torque provided by the hydraulic machine is not generated by an increasing flow of hydraulic fluid. To the contrary, with the minimal passage between the hydraulic input and output ports of the hydraulic machine blocked by a pressure valve, there is no significant volumetric flow of hydraulic fluid at any time. The only flow of fluid is a relatively small blow-by in response to the pressure being developed within the hydraulic machine accompanied by a concomitant replenishing of the blow-by to the low pressure side of the hydraulic machine from the hydraulic system&#39;s conventional charge pump. In effect, the hydraulic machine is operating like a hydraulic “jack”. Each infinitely-variable movement of the swash plate corresponds to the cranking of the jack handle, causing movement of the pistons of the machine to create ever-increasing levels of hydraulic pressure that act as resistance torque to slow the rotation of the swash plate, in a manner similar to the way that each crank of the jack handle increases pressure in the small hydraulic jack to raise the load slowly without any appreciable flow of hydraulic fluid. 
   When vehicle acceleration is desired, the minimal fluid passage that connects the input and output ports of the hydraulic machine is closed off by the pressure valve so that piston movement is limited by the just-mentioned minimal flow of blow-by as pressure is built up by the small angular adjustment of the swash plate that is rapidly rotating at the speed of the control gear. The minimal blow-by, which is less than 5% of the volume of fluid blocked within the machine by the closed valve, permits the angle of the rotating swash plate to be increased, increasing the resistance torque that slows the rotation of the control gear. The minimal blow-by at 1700 rpm is preferably less than 1 gal/min and is conventionally replenished to the low pressure side of the hydraulic machine by a small charge pump. 
   A transmission of the present invention provides a significant gain in engine efficiency by using the just-described simplified hydraulic-jack apparatus rather than the vehicle engine as the primary means for vehicle acceleration. Efficiency is increased during all accelerations up to highway cruising speeds because: (1) the vehicle engine remains at idle speeds or at a slightly increased rpm level within a continuum of relatively low rpm&#39;s (preferably 750-±1500 rpm at a rate of typically 75-100 rpm/sec), and at the same time (2) the simplified transmission provides consecutive infinitely-variable increases in output rpm while concomitantly providing consecutive proportional infinitely-variable decreases in torque (through gear ratio decreases). Because the transmission generates such extremely large starting torques at very low vehicle speeds, and because the changes in torque and vehicle speed remain proportional, the horsepower expended by the engine may thereby be more closely matched to the needs dictated by road and traffic conditions. 
   The transmission hydraulics are only active when providing the infinitely-decreasing gear reduction during the acceleration process. When the vehicle is stopped, the swash plate is returned to 0° and the pressure valve is opened, deactivating the hydraulic-jack effect, and the deactivated hydraulics consume minimal, if any, horsepower. When the vehicle is at cruising speeds, the hydraulics are pressurized and remain locked to hold the control gear in its stopped position, like a load being held up by a hydraulic jack, again consuming minimal, if any, horsepower except for the energy required for the charge pump to replenish the fluid lost in blow-by. 
   In actual vehicle testing, a transmission of the present invention reasonably accelerated the vehicle to 30 mph with a relatively minor increase in engine speed (e.g., 750-1000 rpm at about 75-100 rpm/sec). This is a significant improvement over the relative inefficiency of present conventional transmissions that achieve vehicle acceleration by rapidly increasing engine speed to over 1500 rpm during vehicle acceleration, unnecessarily wasting engine efficiency. In actual in-vehicle testing, a transmission of the present invention was found to consume approximately half as much fuel as conventional automatic when in a stopped condition with the vehicle in “drive” (e.g., stopped at a traffic light). 
   Of course, many operators lack the expertise or patience to learn the manual control procedures just explained above, and many others are unnecessarily “heavy footed” on the accelerator pedal. Therefore, in one embodiment of the present invention vehicle operation is computer-assisted. Such computer programs sense the angle of the accelerator pedal, as well as the rate at which this angle has been increased or decreased by the operator, to progressively select engine speeds from a continuum of relatively low rpm, the rate of engine speed progression being controlled to optimize the horsepower/fuel consumption for the desired acceleration rate indicated by the operator&#39;s actions. After the vehicle reaches a desired speed level, as indicated by the operator&#39;s released angle of the accelerator, the computer backs off the engine to the lowest rpm level necessary to maintain that speed. 
   Recent Torvec IVT transmissions have been much smaller and lighter than the conventional transmissions that they replace, and a transmission of the present invention is even smaller and has significantly less volume weight than the earlier IVT designs, since it omits one complete hydraulic machine. 
   In another embodiment of the present invention, a second hydraulic machine is included to create a hybrid drive. The vehicle&#39;s primary drive is provided by either a gas or diesel engine using the invention&#39;s just-described simplified hydraulic transmission. However, an accumulator assembly is added to the structure (a) to store the kinetic energy of the vehicle during coasting or braking in the form of pressurized hydraulic fluid and (b) to reuse that stored energy to assist in the acceleration or driving of the vehicle. The rotation of the vehicle&#39;s drive shaft during coasting and braking conditions is used as input to a hydraulic machine, acting as a pump, to deliver hydraulic fluid from a reservoir to a pressurized tank. To assist in the acceleration of the vehicle, this same hydraulic machine, acting as a motor, is driven by the stored pressurized fluid to provide supplemental driving torque to the vehicle&#39;s drive shaft. 

   
     BRIEF DESCRIPTION OF THE DRAWINGS 
       FIG. 1  is a schematic view of a first embodiment of the present invention, showing a cross section of the orbital gear complex. 
       FIG. 2  is a block diagram of the transmission in  FIG. 1 , showing in schematic form a cross section of the invention&#39;s hydraulic jack machine. 
       FIG. 3  is a partially schematic view of a second embodiment of the present invention including accumulator apparatus appropriate for use in hybrid vehicles. 
   

   DETAILED DESCRIPTION OF THE INVENTION 
     FIG. 1  and  FIG. 2  are schematic diagrams of a remarkably small and compact transmission in embodiments of the present invention attached to the crankshaft  12  of a primary engine  10  that provides an input for an orbital gear complex  14  which is in combination with a rotary control device that is disclosed in this preferred embodiment as a hydraulic jack machine  16 . An input shaft  18  is splined to engine crankshaft  12 , both of which are aligned along a first axis  13 . A central drive plate  20  is positioned between the two end plates  22 , and these just-named three elements together form the orbital web of the transmission that also rotates about first axis  13 . Input shaft  18  is also splined to central drive plate  20 . End plates  22  support the respective ends of orbit shaft  24  that carries a cluster gear that includes cluster gear  26  and cluster gear  28 . While a preferred orbital gear complex comprises at least two or three sets of orbit shafts  24  and cluster gears  26 / 28 , only one set is shown for clarity. Also, engine crankshaft  12  may alternatively be splined directly to the central drive plate  20 . Central drive plate  20  has openings to provide clearance for cluster gears  26 / 28 , and a control gear  30  meshes with cluster gears  26 , while cluster gears  28  mesh with an output gear  32  coupled to a transmission output shaft  34  that, in turn, is connected to a vehicle drive shaft  35  (as will be explained in further detail below). 
   Control gear  30  is fixed to a control drive gear  36 , and both control gears  30 ,  36  are similarly fixed to a hollow shaft  38  that circumscribes transmission input shaft  18 . Control drive gear  36  is in mesh with a hydraulic drive gear  40  fixed to the drive shaft  42  of hydraulic machine  16  that creates the resistance torque that controls the output of orbital gear complex  14 . Control gear  30  is larger than cluster gear  26 , and cluster gear  28  is larger than output gear  32 . 
   In one preferred embodiment of the present invention, the gear tooth ratios for the orbital gearing are as follows, with reference numerals from  FIG. 1  and  FIG. 2 : 
   
     
       
             
             
             
           
             
             
             
           
         
             
                 
                 
             
             
                 
               Gear 
               No. of Teeth 
             
             
                 
                 
             
           
           
             
                 
             
           
        
         
             
               a. 
               Control gear 30 
               32 
             
             
               b. 
               Cluster gear 26 (in mesh w/30) 
               19 
             
             
               c. 
               Cluster gear 26 (fixed to 26) 
               33 
             
             
               d. 
               Output gear 32 (in mesh w/28) 
               22 
             
             
               e. 
               Control drive gear 36 (fixed to 30) 
               60 
             
             
               f. 
               Hyd. machine drive gear 40 (in mesh w/36) 
               30 
             
             
                 
             
           
        
       
     
   
   Hydraulic jack machine  16 , which operates as the transmission&#39;s rotary control device in a disclosed preferred embodiment, includes a plurality of pistons  44  arranged in cylinders (not individually shown). The stroke of the pistons  44  is controlled by the position of an adjustable swash plate  46  that rotates with drive shaft  42  and hydraulic drive gear  40 . The cylinder block  48  includes a cylinder for each piston, each cylinder having input and output ports  50  connected through only a very minimal passage  52  closable by a fluid pressure valve  54  that also serves as a pressure relief valve (e.g., for avoiding increases in pressure above 4000 psi within machine  16 ). 
   When swash plate  46  is set at 0°, drive shaft  42  and swash plate  46  may freely rotate without resulting in any significant increase of fluid pressure in any portion of hydraulic machine  16 , including minimal passage  52 . However, when pressure valve  54  is closed, blocking off minimal passage  52 , and swash plate  46  is moved in a forward direction, the increasing inclination of swash plate  46  results in increasing hydraulic pressure within the hydraulic machine, slowing the rotation of swash plate  46 , drive shaft  42 , hydraulic drive gear  40  and control drive gear  36 , providing a resistance torque that decreases rotation of control gear  30  proportional to the increase of the resistance torque. This resistance torque varies directly with the fluid pressure in hydraulic machine  16 , and when swash plate  46  is moved to a predetermined maximum angle, the resistance torque prevents rotation of control gear  30 . The changes in hydraulic pressure just described all preferably occur with no fluid motion other than a minimal blow-by replenished by a small conventional charge pump (not shown) to the low pressure side of hydraulic machine  16  at a maximum rate of less than one gallon per minute. 
   In a disclosed preferred embodiment, output shaft  34  from orbital gear complex  14  preferably connects through a standard clutch mechanism  56  to a standard “forward/reverse” gear complex  58 , this gear change being conventionally controlled by a standard gear-shift lever. While the final output of the forward and/or the reverse gearing of complex  58  can remain at 1:1 with the transmission output, some differing output gear ratios may be desired in some designs. Also, a computer  60  preferably monitors (a) the vehicle accelerator pedal  62  (both position and rate of change), (b) a manual shift lever  63 , and (c) hydraulic fluid pressure in hydraulic machine  16  by a fluid pressure sensor  64  to control (d) adjustment of swash plate  46 , (e) operation of clutch  56 , and (f) adjustment of fluid pressure valve  54 . 
   Start-Up 
   When the vehicle is stationary and the engine is first started, the following events preferably occur: The engine begins to operate at idle (e.g., 750 rpm). The orbital web  20 ,  22  of small gear complex  14  rotates with engine crankshaft  12  at engine speed. The wheels of the parked vehicle are standing still on the terrain and, since transmission output gear  32  is connected to the vehicle drive shaft  35 , output gear  32  is held in a stopped condition. With orbital web  20 ,  22  rotating orbit shaft  24  and cluster gears  26 ,  28  about first axis  13  while output gear  32  is held stopped, cluster gear  28  rolls around stopped output gear  32  as the orbital web moves with the engine drive. Under these conditions, with the preferred gear ratios indicated above and with swash plate  46  of hydraulic machine  16  set at 0°, control gear  30  rotates at approximately one-half the engine input speed (e.g., 300 rpm), and hydraulic drive gear  40 , shaft  42 , and swash plate  46  all merely rotate freely at some predetermined overdrive rate faster than the speed of control gear  30 , adding only a minimal frictional load. Once again, special attention is called to the fact that the hydraulic machine disclosed in a preferred embodiment herein uses a variation of the above-mentioned prior art Torvec long-piston hydraulic machine disclosed in U.S. Pat. No. 6,983,680 and U.S. 2004/0168567, which assures the successful operation of the just-described neutral “minimal-load” condition. 
   From Standing Stop 
   Upon vehicle startup from a standing stop, the following events preferably occur: While engine  10  remains at idle (e.g., 750 rpm), pressure valve  54  is closed and swash plate  46  is initially moved in the forward direction, either manually or under computer control in response to the depression of accelerator pedal  62 . The immediate pressure build-up within hydraulic jack machine  16  results in sufficient blow-by to permit swash plate  46  to move about 1-1.5°, and this same immediate pressure increase causes a slow down of control gear  30  from its free-wheeling speed at approximately one-half the idling speed of the engine (e.g., 300 rpm). Gear complex  14  responds to this slow down of control gear  30  by creating a momentary near-infinite gear reduction at the output gear that, in a fraction of a second, drops to 1000-300:1 gear reduction, starting the vehicle&#39;s wheels to turn at very slow rpm with very high torque. 
   Thereafter, the vehicle is accelerated in response to the continued movement of swash plate  46  in the forward direction. However, it is important to note that blow-by in closed hydraulic jack machine  16  remains constant (e.g., less than 5% of the total volume of fluid blocked within the machine by the closed valve  54 ) and that the blow-by determines the maximum rate at which the angle of swash plate  46  can continue to increase. Nonetheless, this maximum rate is relatively fast, and pressure in hydraulic machine  16  increases in direct proportion to the movement of swash plate  46 . This increasing pressure creates resistance torque that opposes and slows the rotation of swash plate  46 , hydraulic machine drive shaft  42 , hydraulic drive gear  40 , control drive gear  36 , and control gear  30 . The increasing slow down of control gear  30  results in the concomitant gradual increase in the rotation of transmission output shaft  34  at the just-described extremely high gear ratio that quickly drops to about 30-20:1, multiplying the engine torque proportionally, starting to move the vehicle wheels. 
   This forward movement of swash plate  46  continues as the vehicle accelerates, further lowering the gear ratio, until the vehicle reaches around 30-40 mph. At this point, the following conditions occur almost simultaneously: (a) swash plate  46  reaches a maximum angle (e.g., 25°); (b) control gear  30  stops; (c) the hydraulic pressure in hydraulic machine  16  remains “locked” (like a hydraulic jack), exerting a constant back pressure that maintains control gear  30  in its stopped condition; and (d) transmission output gear  32  is running at a predetermined overdrive condition as efficiently as if it were held by a clutch. 
   The locked condition of hydraulic machine  16  is maintained as the continuing blow-by (e.g., less than 1 gal/min at vehicle speeds of 50 mph) is conventionally replaced to the low pressure (suction) side of the machine by a small charge pump. 
   When Cruising 
   At highway cruising speeds (i.e., with swash plate  46  at 25° and control gear  30  stopped), when greater drive torque is required, such as for maintaining speed on an incline or passing another vehicle, the operator merely moves shift lever  63  slightly back from its limit position. This is all that is required to move swash plate  46  to a slightly lower angle (e.g., 22°), thereby re-starting movement of hydraulic pistons  44  and control gear  30 , to increase the transmission gear-ratio and output torque. 
   The vehicle may be provided with a well-known “cruise control” feature. If so, when traveling under cruise control at some desired cruising speed and the vehicle encounters a hill, the increased load on the transmission is noted by the operator, or through fluid pressure sensor  64  in minimal passage  52  by computer  60 , and this pressure increase is compensated by moving swash plate  46  back a few degrees (e.g., from 25° to 22°) either by computer input or by manual movement of shift lever  63  back slightly from its maximum (e.g., 25°) position. This causes some reduction of pressure within hydraulic machine  16  that, in turn, results in some movement of control gear  30  to cause an increase in the gear ratio within the transmission, resulting in an increase in output torque until the vehicle again reaches the desired cruising speed and the pressure within the hydraulic system once again becomes balanced. Swash plate  46  is returned to the maximum (e.g., 25°) position during this increase in vehicle speed, increasing resistance torque to once again stop control gear  30 , and the vehicle maintains its desired speed. 
   Similarly, when it is desired to slow the vehicle from a cruising speed, accelerator pedal  62  is released and shift lever  63  is moved back towards the 0° swash plate position, creating increasing braking torque from the slowed engine through the resulting rapidly increasing gear-ratios. Should shift lever  63  near the 0° swash plate position, clutch  56  is engaged before the vehicle&#39;s drive wheels are locked. 
   Special attention is called to the fact that hydraulic machine  16  is not operating like a conventional pump or motor, and thus, the increasing resistance torque provided by hydraulic machine  16  is not generated by an increasing flow of hydraulic fluid. To the contrary, with minimal passage  52  between hydraulic input and output ports  50  blocked by pressure valve  54 , there is no significant volumetric flow of hydraulic fluid at any time. As indicated above, the only flow of fluid is a relatively small blow-by in response to the pressure being developed within hydraulic machine  16  accompanied by a concomitant replenishing of the blow-by to the low pressure side from a conventional charge pump. In effect, hydraulic machine  16  operates like a hydraulic “jack”. Each successive movement of swash plate  46  corresponds to the cranking of the jack handle, causing movement of the pistons of the machine to create ever-increasing levels of hydraulic pressure that act as resistance torque to slow the rotation of swash plate  46 , in a manner similar to the way that each crank of the jack handle increases pressure in the small hydraulic jack to slowly raise the load without any appreciable flow of hydraulic fluid. 
   Special attention is also called to another very important feature of the invention. As indicated above, when the vehicle is stopped and there is no movement of output gear  32 , the orbital gearing creates a mechanical advantage of the engine input to cause control gear  30  to rotate at a predetermined reduction of the idling engine speed. The gear ratio between hydraulic drive gear  42  and control-drive-gear  36 /control gear  30  is intentionally selected to create the same mechanical advantage for the resistance torque pressure developed by the hydraulic machine  16  as that resistance torque enters and affects the orbital gearing and the transmission output. Thus, in effect, the hydraulic resistance torque that slows control gear  30  enters the gear complex at a reduction that matches the engine torque reduction. As just explained above, the preferred embodiment disclosed provides the desired matching-engine resistance torque by selecting a similar 2:1 gear reduction between hydraulic drive gear  40  and control drive gear  36 . However, this reduction can be made even higher to require less initial resistance torque from machine  16  to match engine torque (such as if the transmission is being used with a diesel engine). 
   In actual vehicle testing, a vehicle with a transmission of the present invention readily attained the 30 mph speed while the engine was maintained at a little over 750 rpm. However, the acceleration of the vehicle from stop to this speed may take as long as 12-15 seconds depending on road conditions. Since most operators prefer a faster acceleration, this preference may be achieved manually by no more than a minor increase in the angle of the accelerator pedal. Computer control  60  senses the indicated pedal angle to increase acceleration at a more generally acceptable rate (e.g., 100 rpm/sec). This increased acceleration is achieved without the conventional racing of the engine to over 2000 rpm. Instead, the operator or computer progressively selects relatively low levels of increasing engine rpm, (e.g., from a continuum of 750-1500 rpm). The rate of this engine speed progression is controlled to optimize the horsepower/fuel consumption for the desired acceleration rate, as indicated by the depression angle of the accelerator. After the vehicle reaches a desired speed level, again indicated by accelerator position, the engine speed is backed off to the lowest rpm level necessary to maintain that attained speed. 
   While the following may be a reiteration of the above explanation, some persons may best appreciate the general operation of a transmission of the present invention with the help of the following description of a basic embodiment in which the swash plate of the hydraulic machine is manually controlled by using simple shift lever  63 . After the vehicle is started and while engine  10  is still at idling speed, the standing vehicle is initially accelerated by moving shift lever  63  very slightly in the forward direction, just enough to initiate vehicle movement. Immediately thereafter, accelerator pedal  62  is depressed slightly to increase engine speed by only a few hundred rpm. With only small increments of additional pressure on the accelerator pedal, depending on the rate of vehicle acceleration desired, shift lever  63  is continually moved in the forward direction, until the vehicle reaches a desired speed or until the shift lever  63  reaches its limit (e.g., the 25° swash plate position) and the transmission reaches the predetermined overdrive for sustained cruising operation. The rate of acceleration is controlled completely by the operator, and even the fastest acceptable rates can be achieved with relatively minimal increases in engine rpm. Of course, after a cruising speed appropriate for the terrain and traffic conditions is achieved, the shift lever  63  is allowed to remain in the position that provides the desired cruising speed, and accelerator pedal  62  may be relaxed to a lower level of engine rpm necessary to sustain the attained cruising speed. 
   Accumulator Embodiment 
   Another embodiment of the present invention, shown in partially schematic  FIG. 3 , includes an apparatus that permits operation in regeneration modes similar to those used in well-known “hybrid” vehicle designs. 
   In this embodiment, the transmission converts torque from engine crankshaft  12  to vehicle drive shaft  35  in the manner just explained above with reference to  FIGS. 1 and 2 . A second hydraulic machine  70  is added along with a fluid storage tank  72 , a fluid pressure tank  74 , accumulator transfer gears  76 ,  78 , a clutch  80 , and an accumulator control valve  82 . 
   Whenever the vehicle is braking or coasting, accumulator control valve  82  interconnects hydraulic machine  70  to accumulator tanks  72 ,  74 , and, simultaneously, clutch  80  connects transfer gears  76 ,  78  to the drive shaft of hydraulic machine  70 . During such coasting or braking conditions, the rotation of vehicle drive shaft  35  is increased by gears  76 ,  78  to energize hydraulic machine  70  which acts like a regeneration pump to draw fluid from storage tank  72  and deliver it under pressure to pressure tank  74 . Pressure tank  74  is preferably primarily a steel tube, capped at each end with the interior of pressure tank  74  including a bladder that is filled with a compressible gas in the manner well-known in the art. Regenerative fluid enters pressure tank  74  under pressure that begins to compress the gas in the bladder until pressure tank  74  is full. 
   Storage tank  72  is preferably similar to pressure tank  74  except that it contains no gas-filled bladder and it is initially filled with fluid sufficient for the normal operation of the regeneration system. For many vehicles, elongated tubes that comprise storage tank  72  and pressure tank  74  may be approximately 8′-10′ long and may be positioned along side each of the respective side rails of the vehicle&#39;s frame. 
   During braking or coasting, engine  10  returns to its idle speed and swash plate  46  is readjusted toward the 0° position, causing transmission  14  to produce an ever-increasing reduction and braking torque in output gear  32  and vehicle crankshaft  35 , and clutch  56  is disengaged before the vehicle is braked to a stop, as explained above. 
   As soon as pressure tank  74  is full, or as soon as the vehicle reaches a predetermined minimal operating speed, whichever occurs first, the regeneration circuit is closed off (i.e., valve  82  is moved to its closed position and clutch  80  is disengaged), and the transmission is returned to normal operation (i.e., the swash plates of hydraulic machines  16  and  70  are reoriented to their respective normal positions) based upon the vehicle speed condition then prevailing. 
   When it is desired to restart or reaccelerate the vehicle, hydraulic machine  16  operates in the manner explained above, while clutch  80  is engaged and valve  82  is moved to its open position. The pressurized fluid stored in pressure tank  74  is released to energize hydraulic machine  70  which now acts like a regeneration motor, adding driving torque to engine drive shaft  35  through the reduction of transfer gears  78 ,  76 . During the time that pressurized fluid is being delivered from pressure tank  74 , the regeneration system remains activated (i.e., valve  82  remains open) so that the regeneration fluid is returned to storage tank  72 , while engine  10  remains at idle speed. As soon as the vehicle reaches a desired operating speed, or as soon as pressure tank  74  is depleted of pressurized fluid, whichever occurs first, the regeneration circuit is closed off (i.e., valve  82  is closed and clutch  80  is disengaged), and the speed of engine  10  and the transmission are returned to normal operation. 
   It should be noted that transfer gears  76 ,  78  have an increasing ratio (e.g., 1:3 between engine drive shaft  35  and hydraulic machine  70  when the latter  14  is acting as a pump, increasing the effective speed of hydraulic pump  70  to a multiple (e.g., three times) of the speed of output shaft  35 . Thus, the saved energy from the inertia of output shaft  35  is accumulated in pressure tank  74  at a much faster rate (e.g., three times faster) than it is being lost during the coasting/braking operation. As just indicated above, this stored energy is returned to the vehicle wheels through transfer gears  78 ,  76  by torque-increasing reduction. 
   Accordingly, it is to be understood that the embodiments of the invention herein described are merely illustrative of the application of the principles of the invention. Reference herein to details of the illustrated embodiments is not intended to limit the scope of the claims, which themselves recite those features regarded as essential to the invention.