Patent Publication Number: US-6701709-B2

Title: Cylindrical cam stirling engine drive

Description:
This application claims the benefit of U.S. Provisional Application No. 60/313,309, filed Aug. 18, 2001, which is hereby incorporated by reference herein in its entirety. 
    
    
     BACKGROUND OF THE INVENTION 
     1. Field of the Invention 
     This invention relates broadly to a Stirling engine. More particularly, this invention relates to a cam drive system for a Stirling engine that converts linear mechanical motion of pistons into rotary motion at an output shaft and vice versa. 
     2. State of the Art 
     Stirling engines are heat engines that operate on a closed thermodynamic cycle to convert heat energy into mechanical energy by alternately compressing and expanding a confined working fluid (gas or liquid). As with any heat engine, the engine requires a hot sink and a cold sink and, in the Stirling engine, the confined working fluid is externally heated and cooled. Unlike a steam engine, the working fluid does not change phase at anytime during the thermodynamic cycle. The alternate heating and cooling of the working fluid produces an alternating pressure within the engine. The alternating pressure (or pressure wave) can be converted to mechanical power by several means. For example, the pressure wave can act on pistons, bellows, or diaphragms to convert the pressure wave into mechanical power. Pistons, bellows, and diaphragms produce linear motion that must be converted to rotary motion where rotary motion engine output is desired. 
     There are a number ways to accomplish the conversion of linear motion from the piston into rotary motion. Crankshafts, wobble-plates, swash-plates, cams, and various other means have been used in the past. 
     Theories claim Stirling engine performance can be improved by causing a displacer of the engine to dwell at top dead center and bottom dead center. By dwelling at these positions, the working fluid remains in a heat exchanger of the engine for a longer time resulting in greater energy transfer to or from the walls of the heat exchanger to or from the working fluid. Dwells in motion are relatively easy with cams as compared with other mechanisms such as cranks, wobble-plates, and swash-plates that inherently produce sinusoidal or nearly sinusoidal motion. The cam followers engaging the cams can either be sliding or rolling. 
     High-speed cam design requires attention to the first three derivatives of the displacement function: velocity, acceleration, and jerk. The displacement required is defined by the piston stroke. The shape of the cam curve with respect to rotation is made up of intervals of rise, fall, and dwell. During dwell, there is no piston motion as the cam rotates. The intervals are designed and pieced together so that there preferably are never infinite or excessively high values of acceleration and/or jerk. By controlling acceleration and jerk, the forces on a cam follower and associated moving components can be kept to acceptable levels. This also reduces wear, spalling, and friction on the followers and cam surface in contact with the follower. 
     Except at very low speeds, sliding cam followers require copious lubrication to maintain a hydrodynamic barrier between the follower and cam surface. Lubrication can be achieved by submersion or pump flooding the cam/follower contact area. The follower rides on a thin hydrodynamic layer of lubricant that reduces friction, prevents high speed contact, and carries away heat that may be generated. However, at high speeds, sliding cam followers require a crankcase containing a fluid lubricant such as oil or grease (wet sump). 
     Rolling followers can also be used, but have other problems. U.S. Pat. No. 4,996,953 to Buck describes a grooved cam system for a Stirling engine. When the direction of follower load reverses, as it will with double-acting Siemens-type Stirling pistons, the cam follower alternately contacts both sides of the cam groove as the cam rotates. Because the cam rotates in one direction continuously, the rolling follower must reverse direction instantly when switching contact from an upper surface to a lower surface. This reversing may be acceptable for small light weight follower bearings operating at low speeds but large heavy follower bearings rotating at high speeds have considerable inertia and attempting to instantly reverse direction when contacting the opposite surface results in skidding and destruction of the mating follower and cam surfaces. 
     U.S. Pat. No. 3,385,051 to Kelly teaches a dual blade cam system in which each of two wave-shaped blade cams extends radially outward from the output shaft of the engine. Roller bearings are provided on first and second sides of each of the cams. Blade-type cylindrical cams do not have the problems associated with reversing follower direction of rotation, because for reversing follower loads there are two followers, one above the cam blade, and one below. Each follower is continuously in contact with the same cam surface moving in the same direction. Therefore, there is no skidding. However, these follower assemblies tend to be large, heavy, complex, and expensive. Moreover, unless preloaded, these assemblies can be particularly loud, especially when the load reverses directions and the follower in contact with a cam surface is changed. 
     Some Stirling engines, such as swash-plate drive engines, operate with wet sumps and require sealing at the piston drive rods to prevent oil from entering the working fluid space from the crankcase fluid space as well as containing the working fluid in the working space. Lubricant in the working fluid can contaminate heat exchanger surfaces or plug the fine pores in the regenerator. Contaminated heat exchangers can reduce performance or cause the engine to be inoperable. Contaminated heat exchangers are difficult or impossible to clean. Explosion in the heater can result if the working fluid is air containing oxygen and the contaminating lubricant is flammable. Because of potential contamination or explosion hazard, and the desire to be able to operate in any orientation, dry-sump Stirling engine designs are desirable. 
     SUMMARY OF THE INVENTION 
     It is therefore an object of the invention to provide an improved cam drive mechanism for the conversion of Stirling engine piston linear motion to output shaft rotary motion and vice-versa. 
     It is another object or the invention to provide for optional cam shapes to produce various cam follower (thus piston) motions (displacement, velocity, acceleration, and dwell) such that the Stirling thermodynamics may be exploited by using optimized piston motions. 
     It is also an object of the invention to provide a compact Stirling engine mechanical drive that has low volume and weight with respect to traditional Stirling engines. 
     It is a further object of the invention to provide a high efficiency (low friction) mechanical drive. 
     It is an additional object of the invention to provide a drive mechanism that is easily manufactured and thus less costly to produce. 
     It is yet another object of the invention to provide a drive mechanism that is reliable and has low maintenance requirements. 
     In accord with these objects, which will be discussed in detail below, a Stirling engine is provided having a grooved cam drive mechanism, with cam followers coupled to each piston of the engine and engaged within the grooved cam. Each follower includes a pair of longitudinally displaced bearings. One bearing is adapted to ride along an upper inner surface of the cam, while the other bearing is adapted to ride along a lower inner surface of the cam. 
     More particularly, each follower includes an outer shaft on which a first of the bearings is mounted, and an inner shaft on which a second of the bearings is mounted. A preferably annular space is provided between the inner and outer shafts when the follower is in an unloaded state. Then, when the follower is engaged within the grooved cam, the inner shaft is cantilevered relative to outer shaft within the annular space and results in pre-loading the first bearing against one inner surface of the groove cam and the second bearing against an opposite inner surface of the grooved cam. The pre-loading eliminates excessive noise and increased bearing wear that would otherwise result. 
     In accord with one embodiment of the invention, the axes of rotation for the bearings are offset by a first amount in the unloaded state, and a second lesser amount in the loaded state. 
     In accord with another embodiment, the cam groove has a stepped surface and the bearings of a cam follower have different diameters but a common rotational axis in the unloaded state. When the follower inserted into the groove, the axes of rotation for the bearings are offset, and the larger diameter bearing bears against a surface opposite the step and the smaller diameter bearing bears against a surface of the step. 
     The bearings are preferably crowned, i.e., have a preferably spherically curved surface. This permits line contact with the cam surface thus reduces the effect of the difference in cam surface velocity at different radial distances from the output shaft rotation centerline. Moreover, the tandem pair of bearings on each follower provide greater load carrying capacity. Furthermore, each bearing is dedicated to rotation in only a single direction. 
     The cam and followers of the invention provide an engine capable of operating at high speed and low noise. Furthermore, the cam drive mechanism operates with low wear. Moreover, the cam and followers are easily manufactured, and provide a compact, relatively inexpensive, and light weight assembly. 
     Additional objects and advantages of the invention will become apparent to those skilled in the art upon reference to the detailed description taken in conjunction with the provided figures. 
    
    
     BRIEF DESCRIPTION OF THE DRAWINGS 
     FIG. 1 is a partial cut-away view of a first embodiment of a Stirling engine according to the invention; 
     FIG. 2 is a partial section view of the pistons, cylinders, and heat exchange system of the Stirling engine of the invention; 
     FIG. 3 is a plan elevation of a first embodiment of a cam follower according to the invention; 
     FIG. 4 is a section view along line  4 — 4  in FIG. 3; 
     FIG. 5 is a broken section view of a portion of the cam drive system according to the invention; 
     FIG. 6 is a plan elevation of a second embodiment of a cam follower according to the invention; 
     FIG. 7 is a section view along line  7 — 7  in FIG. 6; 
     FIG. 8 is a broken section view of a portion of a second embodiment of the cam drive system according to the invention, shown having a stepped rectangular groove and the second embodiment of the cam follower; 
     FIG. 9 is a side elevation of an alternative inner shaft for the cam follower of the invention; 
     FIG. 10 is a section view across line  10 — 10  in FIG. 9; 
     FIG. 11 illustrates the dynamic balancing of the cam drive mechanism of the invention; 
     FIG. 12 is a partial cutaway of a second embodiment of a Stirling engine according to the invention; and 
     FIG. 13 is a partial cutaway perspective view of the Stirling engine of FIG.  12 . 
    
    
     DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS 
     Turning now to FIGS. 1 and 2, a first embodiment of a Siemens-type Stirling engine  10  is shown. The engine  10  has four pistons  12 , each provided in a cylinder  14 , and preferably displaced ninety degrees apart. Each piston  12  has a piston seal  16  that prevents passage of a working fluid  17  between a compression space  18  and an expansion space  20  within the cylinder  14 . The pistons  12  are free to move axially in the cylinders  14  and control a cam drive mechanism, described below. While theoretically the engine  10  requires that the positions of the four pistons  12  be maintained in a ninety-degree phase relationship to each other with respect to a rotational axis A r  of a cam  52  and an output shaft  24 , in practice, other phase relationships can be used. 
     In a Siemens-Stirling engine  10 , each cylinder  14  is connected to an adjacent cylinder by a heater  30 , a regenerator  32 , and a cooler  34  (FIG.  2 ). As the pistons  12  translate back and forth in the cylinders  14 , the working fluid  17  is forced to flow in an oscillating fashion to and from the compression spaces  18  and expansion spaces  20  thru the heater  30 , regenerator  32 , cooler  34  and the connecting ducts  36 . 
     Piston rod seals  40  isolate the preferably gaseous working fluid  17  from a gas space  42  in a preferably dry-sump crankcase  46 . Linear piston rod bearings  44  support and locate the piston guide rods  22 . 
     The output shaft  24  is supported in the crankcase  46  by three bearings  48 . An output shaft seal  50  about the shaft  24  contains a preferably pressurized gaseous fluid  51  in the crankcase gas space  42 . A cam  52  is rigidly attached to the output shaft  24 . The cam  52  defines a preferably rectangular groove  54 , with upper and lower surfaces  56 ,  58 . 
     For each piston  12 , a cam follower fitting assembly  60  is provided and includes a mount  68  supporting a cam follower  70  adapted to be inserted into the groove  54 . The mount  68  of the assembly is attached to the lower end of each piston guide rod  22 . A cam follower guide rod  72  is coupled to the bottom of the fitting assembly  60  coaxial with the piston rod  22  and rides in a linear bearing  73 , and a cam follower alignment pin  74  is provided parallel to the follower guide rod  72 . The follower guide rod  72  and alignment pin  74  reciprocate within mating bores  76 ,  78 , respectively, as the piston  12  reciprocates. The follower alignment pin  74  maintains the correct position of the follower assembly  60  with respect to the cam  52  by preventing the follower assembly  60  from rotating about the common axis of the piston guide rod  22  and the follower guide rod  72  due to offset loads on the cam follower  70  which urges the follower  70  away from the cam  52 . 
     Referring to FIGS. 3 and 4, a first embodiment of the cam follower  70  includes first and second longitudinally displaced ball bearings  80 ,  82  that are preferably of equal diameter. The bearings  80 ,  82  are preferably slightly crowned, i.e., have a spherically curved surface. While difficult to see due to the relatively large radius of curvature, this crowning is shown in the figures at  81  and  83 . The crowning permits line contact between the bearings  80 ,  82  and respective inner surfaces  56 ,  58  of the cam groove  54 , and thus reduces the effect of the difference in cam surface velocity at different radial distances from the output shaft rotational axis A r . Moreover, the crowning prevents minor misalignments and deflections from causing binding. 
     The first bearing  80  is mounted on a cylindrical mount  84  of an outer shaft  86 , and the second bearing  82  is mounted on a cylindrical mount  88  of an inner shaft  90  extending through the outer shaft  86 . A centerline of the outer shaft  86  is concentric with a coupling end  92  of the inner shaft  90 . An annular clearance gap  100 , preferably equal all around, is provided between a raised section  102  of the inner shaft  90  and the inner surface of the outer shaft  86 . The cylindrical mount  88  defines a rotational axis A m  that is parallel to but offset by a distance d 1  from a centerline C 1  of the remainder of the inner shaft  90 . Thus, the bearings  80 ,  82  are not rotationally concentric and the outer diameter of bearing  82  is offset from the outer diameter of bearing  80  by a distance d 2  that is equal to distance d 1 . 
     The outer surface of the coupling end  92  the inner shaft  90  includes an outer key slot  93 , and the inner surface of a coupling end  94  of the outer shaft  86  includes an inner key slot  96 . An inner key  98  extends into the slots  93 ,  96  and rotationally locks the inner shaft  90  relative to the outer shaft  86 . 
     Referring to FIG. 5, the follower  70  is coupled within a bore  106  of the mount  68  of the follower assembly  60 . The outer surface of the coupling end  94  of the outer shaft  86  includes a outer key slot  108 , and the mount bore  106  includes an inner key slot  110 . An outer key  112  extends into the key slots  108 ,  110  and rotationally locks the outer shaft  86  within the bore  106 . The coupling end  92  of the inner shaft  90  extends through the bore  106 . The coupling end  92  is provided with threads (not shown), and a washer  114  and nut  116  are secured thereon to lock the follower  70  to the mount  68 . The keys  98 ,  112  ensure that the follower is properly oriented in the mount  68  for the desired orientation of bearing offset d 2 . 
     When the follower  70  is coupled to the mount  68 , it is positioned for insertion into the cam groove  54 . Once in the cam groove  54 , the inner shaft  90  is cantilevered along a resilient beam portion  118  relative to outer shaft  86 . That is, because the centerline C 1  of the follower  70  (FIG. 4) is held perpendicular to the rotational axis A r  of the cam (FIG. 1) and by proper choice of the radial clearance gap  100  and the offset d 2 , offset bearing  82  is forced against lower cam surface  58  and bearing  80  is forced against the upper cam surface  56 . A portion of the distance d 2  and gap  100  is used up in bending the resilient portion  118  of inner shaft  90 . This bending of the resilient portion beam  118  produces a preload that appears as a couple acting at contact points on the cam groove surfaces  56 ,  58 . The couple is counteracted by an opposite couple created by forces from the piston guide rod  22  and cam follower guide rod  72  acting on linear bearings  44  and  73 , respectively. The preloading eliminates excessive noise that would otherwise result and provides for extended bearing life, and more efficient operation. 
     In addition, the pistons  12 , piston guide rods  22 , cam follower mount  68 , cam follower  70 , cam follower guide rod  72 , and cam follower alignment pin  74  comprise a rigid assembly that has a centerline C 2  passing through the center of the contact area of the cam follower  70 ; i.e., between the two bearings  80 ,  82 . By locating the piston rod centerline C 2  through the center of the contact area of the cam follower  70 , the moment about the piston rod centerline C 2  is reduced by providing the shortest moment arm from the piston rod centerline to any point of contact between the cam bearings  80 ,  82  and the cam surfaces  56 ,  58 . 
     Referring to FIGS. 1 and 5, in operation, as each piston  12  is forced up and down by alternating pressure in the cylinder  14  (FIG.  2 ), the engagement of the bearings  80 ,  82  of the cam follower  70  with the cam surfaces  56  and  58  force the cam  52  and consequently the output shaft  24  to rotate about rotational axis A r . The cam follower alignment pin  74  slides in its bore  78  in the crankcase  46  and prevents the cam follower assembly  60  from rotating about the axis defined by rods  22  and  72 . 
     In view of the above arrangement for the cam drive mechanism, and assuming a preferred set of parameters in which: 
     i) the spring rate of the cantilevered beam portion  118  of the inner shaft  90  measured at the bearing-to-cam contact point equals 10,000 lbs/inch, 
     ii) the cam groove width=bearing diameter+0.020 inch, and 
     iii) the annular gap  100  between the raised section  102  of the inner shaft  90  and the outer shaft  86 =0.020 inch, Table 1 sets forth various preferred exemplar contact forces created and gaps defined between identified elements during operation and otherwise. 
     
       
         
           
               
             
               
                 TABLE 1 
               
             
            
               
                   
               
               
                 Contact Forces and Gaps for Various Cam and Follower 
               
               
                 Configurations 
               
            
           
           
               
               
               
               
               
               
               
               
               
               
               
               
            
               
                 ID 
                 Configuration 
                 F80a 
                 F82a 
                 F80b 
                 F82b 
                 G80b 
                 G82b 
                 G82a 
                 G80a 
                 G100a 
                 G100b 
               
               
                   
               
               
                 1 
                 Prior to 
                 0 
                 0 
                 0 
                 0 
                 — 
                 — 
                 — 
                 — 
                 0.020 
                 0.020 
               
               
                   
                 Insertion 
               
               
                   
                 into Groove 
               
               
                 2 
                 Inserted into 
                 100 
                 100 
                 0 
                 0 
                 0.020 
                 0.020 
                 0.000 
                 0.000 
                 0.010 
                 0.030 
               
               
                   
                 Groove 
               
               
                 3 
                 Piston Force 
                 200 
                 100 
                 0 
                 0 
                 0.020 
                 0.020 
                 0.000 
                 0.000 
                 0.010 
                 0.030 
               
               
                   
                 Up = 100 
               
               
                 4 
                 Piston Force 
                 1,100 
                 100 
                 0 
                 0 
                 0.020 
                 0.020 
                 0.000 
                 0.000 
                 0.010 
                 0.030 
               
               
                   
                 up = 1,000 
               
               
                 5 
                 Piston Force 
                 0 
                 200 
                 0 
                 0 
                 0.010 
                 0.020 
                 0.000 
                 0.010 
                 0.000 
                 0.040 
               
               
                   
                 Down = 100 
               
               
                 6 
                 Piston Force 
                 0 
                 1,100 
                 0 
                 0 
                 0.010 
                 0.020 
                 0.000 
                 0.010 
                 0.000 
                 0.040 
               
               
                   
                 Down = 1,000 
               
               
                   
               
            
           
         
       
     
     In Table 1, F 80   a  and F 82   a  refer to forces at the surfaces of respective bearings  80 ,  82  which are in contact with respective cam surfaces  56 ,  58 , and F 80   b , F 82   b  refer to forces at a diametric location on bearings  80 ,  82 , respectively. Referring to Table 1 and FIG. 5, G 80   a  refers to the gap or space between bearing  80  and the upper cam surface  56 , and G 80   b  refers to the gap between bearing  80  and the lower cam surface  58 . Likewise, G 82   a  refers to the gap or space between bearing  82  and the lower cam surface  58 , and G 82   b  refers to the gap between bearing  82  and the upper cam surface  56 . Finally, Referring to Table 1 and FIGS. 4 and 5, G 100   a  refers to the gap space  100  between an upper side of the raised section  102  of the inner shaft  90  and the outer shaft  86 , and G 100   b  refers to the gap between a lower side of the raised section of the inner shaft and the outer shaft. 
     Therefore, as indicated at row ID 1  of Table 1, prior to installation of the follower bearings  80 ,  82  into the groove  54 , all contact forces equal 0 lbs. Moreover, gaps G 80   a ,  80   b ,  82   a , and  82   b  are undefined as there is no mating cam surface relative to which a measurement can be made. In addition, there is a uniform annular gap space at 100 between the inner and outer shafts, thereby making gaps G 100   a  and G 100   b  equal. 
     Once the follower is inserted into the groove (row ID 2 ), the contact surfaces of each of the bearings  80 ,  82  is subject to a preloading force F 80   a , F 82   a  of 100 lbs, while gaps G 80   b  and G 82   b  are 0.020 inch, as the cam groove is 0.020 inch wider than the bearing diameters. Gaps G 80   a  and G 82   a  are 0.000 inch, as these are now contact points. Gap G 100   a  is reduced to 0.010 inch, while gap G 100   b  is increased to 0.030 inch because the beam  118  is deflected upward by bearing  82  contacting cam surface  58 . 
     Then, at row ID 3 , when an upward piston force of 100 lbs is added to the follower  70 , forces F 80   b  and F 82   b  are 0, as there is no contact with the cam surfaces at the respective bearing surfaces. Force F 82   a  remains at 100 lbs because the beam  118  has not deflected any more or less, while force F 80   a  equals 200 lbs (the preload of 100 lbs plus the upward piston force of 100 lbs). Gaps G 80   b  and G 82   b  equal 0.020 inch because the bearings  80  and  82  both remain in contact with their respective bearing surfaces, and gaps G 80   a  and G 80   b  consequently remain at 0.000 inch. Gap G 100   a  remains at 0.010 inch and G 100   b  remains at 0.030 inch because there is no relative movement between the inner and outer shafts  86 ,  90 . 
     At row ID 4 , the upward piston force is increased to 1000 lbs. The forces F 80   b , F 82   b  remain at 0. Force F 82   a  remains at 100 lbs because the deflection of the beam is not altered. Force F 80   a  is now at 1100 lbs (the sum of the preload and the upward piston force). The gaps are all as discussed above in row ID 3 . 
     At row ID 5 , a 100 lbs downward force is applied to the piston  12 , and hence the follower  70 . Forces F 80   b  and F 82   b  remain at 0. Force F 82   a  is 200 lbs (the sum of the preload and the piston force), while F 80   a  is 0 because the beam  118  has been deflected by the added 100 lbs force. Gap G 82   b  is 0.020 inch and gap G 82   a  is 0.000 inch because bearing  82  is still in contact with the lower cam surface  56 . Gaps G 80   a  and G 80   b  are each 0.010 inch because the beam has been deflected to a maximum extent. In addition, due to beam deflection, the annular space  100  is converted into a space that is not continuous about the inner shaft, as the inner shaft contacts the outer shaft (FIG.  5 ), making gap G 100   a  equal to 0.000 and gap G 100   b  equal to 0.040 inch. 
     Finally, at row ID 6 , the downward force is increased to 1000 lbs. The forces F 80   b  and F 82   b  remain at 0. Force F 82   a  is at 1100 lbs, while force F 80   a  remains at 0 lbs due to beam deflection. The gaps are all as discussed above with respect to row ID 5 . 
     As such, Table 1 shows that whenever the follower assembly  70  is installed into the groove  54 , there is always a clearance gap at G 80   b  and G 82   b , and forces F 80   b  and F 82   b  are always 0 lbs. Gap G 82   a  is always 0.00 inch, and force F 82   a  is always greater than 0; thus, bearing  82  is always preloaded. There is a condition when the piston force is downward that force F 80   a  equals 0 lb and gap G 80   b  equals 0.010 inch. At this time bearing  80  is not preloaded, but this is only for a portion of the cam revolution. Importantly, both bearings  80 ,  82  revolve in the same direction continuously. 
     The inner shaft  90 , outer shaft  86 , and mating components are easily manufactured, comprise a more compact, light-weight assembly, and should be less expensive than the followers required for bladed cam mechanisms. Moreover, the forces on the follower and associated moving components can be kept to acceptable levels, reducing wear, spalling, and friction on the followers and cam surfaces in contact with the followers. 
     Turning now to FIGS. 6 and 7, a second embodiment of a cam follower  270 , substantially similar to the first embodiment, (with like elements having reference numerals incremented by  200  relative to cam follower  70 ) is shown. The cam follower  270  includes two inline bearings  280 ,  282 . Bearing  280  is mounting on a bearing mount  284  at an end of outer shaft  286 , while bearing  282  is mounted on a bearing mount  288  at an end of inner shaft  290 . Bearing  282  is smaller in diameter than bearing  280 . Unlike inner shaft  90 , all cylindrical surfaces on inner shaft  290  are concentric and thus the bearings  280  and  282  are concentric about centerline C 3  in the free unloaded state. As such, annular gap  100  is equal all around. 
     Referring to FIG. 8, the cam groove  254  includes a step  259 , e.g., on the lower cam surface  258 . The distance between cam surfaces  256 ,  258 , the height of step  259 , and the diameters of bearings  280  and  282  determine the preload when installed. More particularly, step  259  forces bearing  282  out of concentricity with bearing  280 . The cantilever beam section  318  (FIG. 7) of inner shaft  290  is thereby bent, thus producing a preload. 
     Turning now to FIG. 9, an alternate inner shaft  390  is shown which may be substituted for inner shafts  90  and  290  where a lower spring rate of a cantilever beam section  418  may be desired. Portions of the cantilever beam section  418  are removed to reduce the cross-sectional area of the section (FIG.  10 ). This results in a beam that is relatively stiffer in one direction than the other so that deflections in different directions can be controlled. 
     Referring now to FIG. 11, the dynamic balancing of the cam drive mechanism is shown. The rotating cam with its asymmetric mass distribution creates a couple D y ×F x  about the origin O of the x-y coordinate system shown. The moving piston/follower masses create the opposite couple D y ×F y . By correct choice of masses and separation distances, the opposite couples can be made equal and to cancel each other thus dynamically balancing the mechanism. 
     Turning now to FIGS. 12 and 13, a second embodiment of a Stirling engine  410 , substantially similar to the first embodiment (with like elements having reference numerals incremented by  400 ), is shown. The engine  410  includes a dual guide rod design for the cam follower assembly. More particularly, first and second guide rods  476 ,  477  are spaced apart and rigidly attached to the upper and lower portions  447 ,  449  of the crankcase  446 . The guide rods  476 ,  477  extend parallel to the piston rod  422 . The mount  468  of the cam follower assembly  460  includes upper ears  520 ,  521  and lower ears  522 ,  523 , each defining a bore provided with a bearing  524 . Guide rod  476  extends through bearings  524  in ears  520 ,  522 , and guide rod  477  extends through bearings  524  in ears  521 ,  523 . This design provides more rigid guidance to the follower assembly  460  (relative to the single guide rod  76  of the first embodiment). Moreover, this cam follower assembly is significantly shorter than the first embodiment, thereby permitting the overall height of the crankcase  446  to be reduced. Compare the height of crankcase  46  (FIG. 1) with the height of crankcase  446 . Thus, weight and size reduction result. 
     There have been described and illustrated herein several embodiments of a Stirling engine and cam drive mechanism suitable for a Stirling engine. While particular embodiments of the invention have been described, it is not intended that the invention be limited thereto, as it is intended that the invention be as broad in scope as the art will allow and that the specification be read likewise. Thus, while ball bearings have been disclosed for use with the follower, it will be appreciated that other bearings, such as roller and needle bearings, can be used as well. In addition, while a preload of approximately 100 lbs (e.g., 75 lbs to 125 lbs) is preferred, it is recognized that the system can be designed to subject the bearings to other preload forces. Also, where particular gap dimensions have been provided, it is understood that other gap dimensions can be used. Furthermore, while a Siemens-type engine with four pistons/cylinders has been shown, it is understood that other types of Stirling engines with other numbers of pistons and cylinders can be used. Moreover, additional piston sets can be added by adding more grooves displaced axially along the output shaft. Additional piston sets can also operate in the same groove and face axially in the same direction as the original pistons or face in the opposite direction. For example, eight pistons can operate in one groove and maintain a ninety degree phase relationship by using a groove with two cycles per revolution instead of one cycle as shown in FIG.  1 . Also, it is appreciated that the engine can be used as a refrigerator or heat pump, in which rotation of the shaft  54  with attached cam  52  causes the followers  70  to move in the cam groove  54  in a manner that causes the pistons  12  to translate within their respective cylinders  14 . Furthermore, while the preferred description has included pistons within the cylinders, it is understood that bellows, diaphragms, and other mechanisms can be used, and for purposes of simplicity, the term ‘piston’ should be read to include bellows, diaphragms, and such other mechanisms, particularly with respect to the claims. It will therefore be appreciated by those skilled in the art that yet other modifications could be made to the provided invention without deviating from its spirit and scope as claimed.