Abstract:
A system for supporting lateral loads on a piston undergoing reciprocating motion along a longitudinal axis in a cylinder includes a guide link for coupling the piston to a crankshaft undergoing rotary motion about a rotation axis of the crankshaft where the longitudinal axis and the rotation axis are substantially orthogonal to each other. A first guide element is located along the length of the guide link and includes a spring mechanism for urging the first guide element into contact with the guide link. The spring mechanism includes a first spring with a first natural frequency of oscillation and a second spring with a second natural frequency of oscillation. A second guide element is in opposition to the first guide element.

Description:
PRIORITY 
     The present application is a continuation-in-part of U.S. patent application Ser. No. 09/335,392, filed Jun. 17, 1999, which is herein incorporated by reference. 
    
    
     TECHNICAL FIELD 
     The present invention pertains to improvements to an engine and more particularly to improvements relating to mechanical components of a Stirling cycle heat engine or refrigerator which contribute to increased engine operating efficiency and lifetime. 
     BACKGROUND OF THE INVENTION 
     Stirling cycle machines, including engines and refrigerators, have a long technological heritage, described in detail in Walker,  Stirling Engines , Oxford University Press (1980), herein incorporated by reference. The principle underlying the Stirling cycle engine is the mechanical realization of the Stirling thermodynamic cycle: isovolumetric heating of a gas within a cylinder, isothermal expansion of the gas (during which work is performed by driving a piston), isovolumetric cooling, and isothermal compression. The Stirling cycle refrigerator is also the mechanical realization of a thermodynamic cycle which approximates the ideal Stirling thermodynamic cycle. In an ideal Stirling thermodynamic cycle, the working fluid undergoes successive cycles of isovolumetric heating, isothermal expansion, isovolumetric cooling and isothermal compression. Practical realizations of the cycle, wherein the stages are neither isovolumetric nor isothermal, are within the scope of the present invention and may be referred to within the present description in the language of the ideal case without limitation of the scope of the invention as claimed. Various aspects of the present invention apply to both Stirling cycle engines and Stirling cycle refrigerators, which are referred to collectively as Stirling cycle machines in the present description and in any appended claims. 
     The principle of operation of a Stirling engine is readily described with reference to FIGS. 1 a - 1   e , wherein identical numerals are used to identify the same or similar parts. Many mechanical layouts of Stirling cycle machines are known in the art, and the particular Stirling engine designated generally by numeral  10  is shown merely for illustrative purposes. In FIGS. 1 a  to  1   d , piston  12  and a displacer  14  move in phased reciprocating motion within cylinders  16  which, in some embodiments of the Stirling engine, may be a single cylinder. Typically, a displacer  14  does not have a seal. However, a displacer  14  with a seal (commonly known as an expansion piston) may be used. Both a displacer without a seal or an expansion piston will work in a Stirling engine in an “expansion” cylinder. A working fluid contained within cylinders  16  is constrained by seals from escaping around piston  12  and displacer  14 . The working fluid is chosen for its thermodynamic properties, as discussed in the description below, and is typically helium at a pressure of several atmospheres. The position of displacer  14  governs whether the working fluid is in contact with hot interface  18  or cold interface  20 , corresponding, respectively, to the interfaces at which heat is supplied to and extracted from the working fluid. The supply and extraction of heat is discussed in further detail below. The volume of working fluid governed by the position of the piston  12  is referred to as compression space  22 . 
     During the first phase of the engine cycle, the starting condition of which is depicted in FIG. 1 a , piston  12  compresses the fluid in compression space  22 . The compression occurs at a substantially constant temperature because heat is extracted from the fluid to the ambient environment. In practice, a cooler (not shown) is provided. The condition of engine  10  after compression is depicted in FIG. 1 b . During the second phase of the cycle, displacer  14  moves in the direction of cold interface  20 , with the working fluid displaced from the region of cold interface  20  to the region of hot interface  18 . This phase may be referred to as the transfer phase. At the end of the transfer phase, the fluid is at a higher pressure since the working fluid has been heated at constant volume. The increased pressure is depicted symbolically in FIG. 1 c  by the reading of pressure gauge  24 . 
     During the third phase (the expansion stroke) of the engine cycle, the volume of compression space  22  increases as heat is drawn in from outside engine  10 , thereby converting heat to work. In practice, heat is provided to the fluid by means of a heater (not shown). At the end of the expansion phase, compression space  22  is full of cold fluid, as depicted in FIG. 1 d . During the fourth phase of the engine cycle, fluid is transferred from the region of hot interface  18  to the region of cold interface  20  by motion of displacer  14  in the opposing sense. At the end of this second transfer phase, the fluid fills compression space  22  and cold interface  20 , as depicted in FIG. 1 a , and is ready for a repetition of the compression phase. The Stirling cycle is depicted in a P-V (pressure-volume) diagram as shown in FIG. 1 e.    
     Additionally, on passing from the region of hot interface  18  to the region of cold interface  20 , the fluid may pass through a regenerator (not shown). The regenerator may be a matrix of material having a large ratio of surface area to volume which serves to absorb heat from the fluid when it enters hot from the region of hot interface  18  and to heat the fluid when it passes from the region of cold interface  20 . 
     The principle of operation of a Stirling cycle refrigerator can also be described with reference to FIGS. 1 a - 1   e , wherein identical numerals are used to identify the same or similar parts. The differences between the engine described above and a Stirling machine employed as a refrigerator are that compression volume  22  is typically in thermal communication with ambient temperature and expansion volume  24  is connected to an external cooling load (not shown). Refrigerator operation requires net work input. 
     Stirling cycle engines have not generally been used in practical applications, and Stirling cycle refrigerators have been limited to the specialty field of cryogenics, due to several daunting engineering challenges to their development. These involve such practical considerations as efficiency, vibration, lifetime, and cost. The instant invention addresses these considerations. 
     A major problem encountered in the design of certain engines, including the compact Stirling engine, is that of the friction generated by a sliding piston resulting from misalignment of the piston in the cylinder and lateral forces exerted on the piston by the linkage of the piston to a rotating crankshaft. In a typical prior art piston-crankshaft configuration such as that depicted in FIG. 2, a piston  10  executes reciprocating motion along longitudinal direction  12  within cylinder  14 . Piston  10  is coupled to an end of connecting rod  16  at a pivot such as a pin  18 . The other end  20  of connecting rod  16  is coupled to a crankshaft  22  at a fixed distance  24  from the axis of rotation  26  of the crankshaft. As crankshaft  22  rotates about the axis of rotation  26 , the connecting rod end  20  connected to the crankshaft traces a circular path while the connecting rod end  28  connected to the piston  10  traces a linear path  30 . The connecting rod angle  32 , defined by the connecting rod longitudinal axis  34  and the axis  30  of the piston, will vary as the crankshaft rotates. The maximum connecting rod angle will depend on the connecting rod offset on the crankshaft and on the length of the connecting rod. The force transmitted by the connecting rod may be decomposed into a longitudinal component  38  and a lateral component  40 , each acting through pin  18  on piston  10 . Minimizing the maximum connecting rod angle  32  will decrease the lateral forces  40  on the piston and thereby reduce friction and increase the mechanical efficiency of the engine. The maximum connecting rod angle can be minimized by decreasing the connecting rod offset  24  on the crankshaft  22  or by increasing the connecting rod length. However, decreasing the connecting rod offset on the crankshaft will decrease the stroke length of the piston and result in less Δ (pV) work per piston cycle. Increasing the connecting rod length can not reduce the connecting rod angle to zero but does increase the size of the crankcase resulting in a less portable and compact engine. 
     Referring now to the prior art engine configuration of FIG. 3, it is known that in order to reduce the lateral forces on the piston, a guide link  42  may be used as a guidance system to take up lateral forces while keeping the motion of piston  10  constrained to linear motion. In a guide link design, the connecting rod  16  is replaced by the combination of guide link  42  and a connecting rod  16 . Guide link  42  is aligned with the wall  44  of piston cylinder  14  and is constrained to follow linear motion by two sets of rollers or guides, forward rollers  46  and rear rollers  48 . The end  50  of guide link  42  is connected to connecting rod  16  which is, in turn, connected to crankshaft  22  at a distance offset from the rotational axis  26  of the crankshaft. Guide link  42  acts as an extension of piston  10  and the lateral forces on the piston that would normally be transmitted to cylinder walls  44  are instead taken up by the two sets of rollers  46  and  48 . Both sets of rollers  46  and  48  are required to maintain the alignment of guide link  42  and to take up the lateral forces being transmitted to the guide link by the connecting rod. The distance d between the forward set of rollers and the rear set of rollers may be reduced to decrease the size of the crankcase (not shown). However, reducing the distance between the rollers will increase the lateral load  54  on the forward set of rollers since the rear roller set acts as a fulcrum  56  to a lever  58  defined by the connection point  52  of the guide link and connecting rod  16 . 
     The guide link will generally increase the size of the crankcase because the guide link must be of sufficient length that when the piston is at its maximum extension into the piston cylinder, the guide link extends beyond the piston cylinder so that the two sets of rollers maintain contact and alignment with the guide link. 
     SUMMARY OF THE INVENTION 
     In accordance with one aspect of the invention, a system for supporting lateral loads on a piston undergoing reciprocating motion along a longitudinal axis in a cylinder includes a guide link coupling the piston to a crankshaft undergoing rotary motion about a rotation axis of the crankshaft. A first guide element is located along the length of the guide link and includes a spring mechanism for urging the first guide element into contact with the guide link. The spring mechanism includes a first spring with a first natural frequency of oscillation and a second spring with a second natural frequency of oscillation. A second guide element is in opposition to the first guide element. In one embodiment, the first guide element is a roller having a rim in rolling contact with the guide link and the second guide element is a roller with a rim in rolling contact with the guide link. 
     In a further embodiment, the second guide element includes a precision positioner for positioning the second guide element with respect to the longitudinal axis. The precision positioner may be a vernier mechanism having an eccentric shaft for varying a distance between the second guide element and the longitudinal axis. 
     In accordance with another aspect of the invention, a linkage for coupling a piston undergoing reciprocating linear motion along a longitudinal axis to a crankshaft undergoing rotary motion about a rotation axis of the crankshaft includes a guide link having a first end proximal to the piston and coupled to the piston and a second end distal to the piston such that the rotation axis is disposed between the proximal end and the distal end of the guide link. A connecting rod is rotably connected to the end of the guide link distal to the piston at a rod connection point at a connecting end of the connecting rod. The connecting rod is coupled to the crankshaft at a crankshaft connection point on a crankshaft end of the connecting rod, where the crankshaft connection point is offset from the rotation axis of the crankshaft. A guide link guide assembly supports lateral loads at the distal end of the guide link and includes a first roller having a center of rotation fixed with respect to the rotation axis of the crankshaft and a rim in rolling contact with the distal end of the guide link. A spring mechanism is used to urge the rim of the first roller into contact with the distal end of the guide link. The spring mechanism includes a first spring with a first natural frequency of oscillation and a second spring with a second natural frequency of oscillation. 
     In one embodiment, the guide link guide assembly further includes a second roller in opposition to the first roller and having a center of rotation and a rim in rolling contact with the distal end of the piston. The second roller may include a precision positioner to position the center of rotation of the second roller with respect to the longitudinal axis. In a further embodiment, the precision positioner is a vernier mechanism having an eccentric shaft for varying the distance between the center of rotation of the second roller and the longitudinal axis. 
     In accordance with yet another aspect of the invention, an improvement is provided to a Stirling cycle machine of the type where at least one piston undergoes reciprocating motion along a longitudinal axis in a cylinder. The piston is coupled to a crankshaft undergoing rotary motion about a rotation axis using a guide link having a first end proximal to the piston and coupled to the piston and a second end distal to the piston. The improvement has a guide link guide assembly including a spring mechanism for urging the rim of a first roller into contact with the distal end of the guide link where the spring mechanism includes a first spring with a first natural frequency of oscillation and a second spring with a second natural frequency of oscillation. 
    
    
     BRIEF DESCRIPTION OF THE DRAWINGS 
     The invention will be more readily understood by reference to the following description, taken with the accompanying drawings, in which: 
     FIGS  1   a - 1   e  depict the principle of operation of a prior art Stirling cycle machine. 
     FIG. 2 is a cross-sectional view of a prior art linkage for an engine. 
     FIG. 3 is a cross-sectional view of a second prior art linkage for an engine, the linkage having a guide link. 
     FIG. 4 is a cross-sectional view of a folded guide link linkage for an engine in accordance with a preferred embodiment of the present invention. 
     FIG. 5 is a perspective view of a guide link and guide wheel assembly in accordance with an embodiment of the invention. 
     FIG. 6 a  is a cross-sectional view of a piston and guide assembly for allowing the precision alignment of piston motion using vernier alignment in accordance with a preferred embodiment of the invention. 
     FIG. 6 b  is a side view of the precision alignment mechanism in accordance with an embodiment of the invention. 
     FIG. 6 c  is a perspective view of the precision alignment mechanism of FIG. 6 b  in accordance with an embodiment of the invention. 
     FIG. 6 d  is a top view of the precision alignment mechanism of FIG. 6 b  in accordance with an embodiment of the invention. 
     FIG. 6 e  is a top view of the precision alignment mechanism of FIG. 6 b  with both the locking holes and the bracket holes showing in accordance with an embodiment of the invention. 
     FIG. 7 is a cross-sectional view of a folded guide link linkage for a two-piston machine such as a Stirling cycle machine in accordance with a preferred embodiment of the present invention. 
     FIG. 8 is a perspective view of one embodiment of the dual folded guide link linkage of FIG.  7 . 
    
    
     DETAILED DESCRIPTION OF PREFERRED EMBODIMENTS 
     Referring now to FIG. 4, a schematic diagram is shown of a folded guide link linkage designated generally by numeral  400 . A piston  401  is rigidly coupled to the piston end of a guide link  403  at a piston connection point  402 . Guide link  403  is rotatably connected to a connecting rod  405  at a rod connection point  404 . The piston connection point  402  and the rod connection point  404  define the longitudinal axis  420  of guide link  403 . 
     Connecting rod  405  is rotatably connected to a crankshaft  406  at a crankshaft connection point  408  which is offset a fixed distance from the crankshaft axis of rotation  407 . The crankshaft axis of rotation  407  is orthogonal to the longitudinal axis  420  of the guide link  403  and the crankshaft axis of rotation  407  is disposed between the rod connection point  404  and the piston connection point  402 . In a preferred embodiment, the crankshaft axis of rotation  407  intersects the longitudinal axis  420 . 
     An end  414  of guide link  403  is constrained between a first roller  409  and an opposing second roller  411 . The centers of roller  409  and roller  411  are designated respectively by numerals  410  and  412 . The position of guide link piston linkage  400  depicted in FIG. 4 is that of mid-stroke point in the cycle. This occurs when the radius  416  between the crankshaft connection point  408  and the crankshaft axis of rotation  407  is orthogonal to the plane defined by the crankshaft axis of rotation  407  and the longitudinal axis of the guide link  403 . In a preferred embodiment, the rollers  409 ,  411  are placed with respect to the guide link  403  in such a manner that the rod connection point  404  is in the line defined by the centers  410 ,  412  of the rollers  409 ,  411  at mid-stroke. As rollers  409 ,  411  wear during use, the misalignment of the guide link will increase. In a preferred embodiment, the first roller  409  is spring loaded to maintain rolling contact with the guide link  403 . In accordance with embodiments of the invention, guide link  403  may comprise subcomponents such that the portion  413  of the guide link proximal to the piston may be a lightweight material such as aluminum, whereas the “tail” portion  414  of the guide link distal to the piston may be a durable material such as steel to reduce wear due to friction at rollers  409  and  411 . 
     Alignment of the longitudinal axis  420  of the guide link  403  with respect to piston cylinder  14  is maintained by the rollers  409 ,  411  and by the piston  401 . As crankshaft  406  rotates about the crankshaft axis of rotation  407 , the rod connection point  404  traces a linear path along the longitudinal axis  420  of the guide link  403 . Piston  401  and guide link  403  form a lever with the piston  401  at one end of the lever and the rod end  414  of the guide link  403  at the other end of the lever. The fulcrum of the lever is on the line defined by the centers  410 ,  412  of the rollers  409 ,  411 . The lever is loaded by a force applied at the rod connection point  404 . As rod connection point  404  traces a path along the longitudinal axis of the guide link  403 , the distance between the rod connection point  404  and the fulcrum, the first lever arm, will vary from zero to one-half the stroke distance of the piston  401 . The second lever arm is the distance from the fulcrum to the piston  401 . The lever ratio of the second lever arm to the first lever arm will always be greater than one, preferably in the range from 5 to 15. The lateral force at the piston  401  will be the forced applied at the rod connection point  404  scaled by the lever ratio; the larger the lever ratio, the smaller the lateral force at the piston  401 . 
     By moving the connection point to the side of the crankshaft axis distal to that of the piston, the distance between the crankshaft axis and the piston cylinder does not have to be increased to accommodate the roller housing. Additionally, only one set of rollers is required for aligning the piston, thereby advantageously reducing the size of the roller housing and the overall size of the engine. In accordance with the invention, while the piston experiences a non-zero lateral force (unlike a standard guide link design where the lateral force of a perfectly aligned piston is zero), the lateral force can be at least an order of magnitude less than that experienced by a simple connecting rod crankshaft arrangement due to the large lever arm created by the guide link. 
     Lateral forces on a piston can give rise to noise and to wear. As mentioned above, roller  409  and roller  411  are used to align the piston  401  and to take up lateral forces being transmitted to the guide link  403  by the connecting rod  405 . Preferably, one of the rollers  409  is spring loaded to maintain rolling contact with the guide link  403 . At least one spring may be used to force the roller  409  (otherwise referred to herein as a guide wheel) against the guide link  403  surface. During operation of an engine, the guide wheel  409  and spring mechanism will typically reciprocate or bounce on the surface of the guide link  403  at or near the natural resonant frequency of the guide wheel and spring combination. This oscillation may result in significant fluctuations in the force supporting the guide link  403  as well as intermittent contact between the guide link  403  and the guide wheel  409 . This, in turn, results in excessive noise, increased wear and decreased efficiency and power output. 
     FIG. 5 is a perspective view of a guide link and guide wheel assembly in accordance with an embodiment of the invention. In FIG. 5, a guide link  500  is supported at its free end by a fixed guide wheel  501  and a spring loaded guide wheel assembly  502 . The guide wheel assembly  502  includes two springs  504 ,  505  and a guide wheel  506 . Springs  504  and  505  force the guide wheel  506  against the guide link  500 . Springs  504  and  505  have the combined force necessary to hold the guide wheel assembly  502  in contact with guide link  500 . In addition, spring  504  and spring  505  each have a different natural frequency of oscillation (i.e., each has a different spring rate). By selecting springs with non-overlapping natural frequencies, at least one spring will advantageously not be in resonance at all times during operation. As mentioned above, the guide wheel assembly  502  will typically reciprocate on the surface of the guide link  500  at or near the natural resonant frequency of the guide wheel and springs. By using two springs with different natural frequencies of oscillation, the resonance of the guide wheel assembly  502  should be eliminated since at least one spring will not be in resonance. 
     Additional friction may be generated by the misalignment of the piston in the cylinder. A solution to the alignment problem is now discussed with reference to FIGS. 6 a - 6   e . FIG. 6 a  shows a schematic diagram of a piston  601  and a guide assembly  609  for allowing precision alignment of piston motion using vernier alignment in accordance with a preferred embodiment of the invention. The piston  601  executes a reciprocating motion along a longitudinal axis  602  in cylinder  600 . A guide link  604  is coupled to the piston  601 . An end of the guide link  604  is constrained between a first roller  605  and an opposing second roller  607 . The centers of roller  605  and roller  607  are designated respectively by numerals  606  and  608 . A piston guide ring  603  may be used at one end of the piston  601  to prevent piston  601  from touching the cylinder  600 . However, if piston  601  is not aligned to move in a straight line along longitudinal axis  602 , it is possible other points along the length of piston  601  not coupled to the guide ring may contact the cylinder  600 . In a preferred embodiment, piston  601  is aligned using rollers  605  and  607  and guide link  604  such that piston  601  moves along the longitudinal axis  602  in a straight line and is substantially centered with respect to cylinder  600 . 
     In accordance with a preferred embodiment of the invention, the piston  601  may be aligned with respect to the piston cylinder  600  by adjusting the position of the center  608  of the second roller  607 . The first roller  605  is spring loaded to maintain rolling contact with the guide link  604 . The second roller  607  is mounted on an eccentric flange such that rotation of the flange causes the second roller  607  to move laterally with respect to longitudinal axis  602 . A single pin (not shown) may be used to secure the second roller  607  into a position. The movement of the second roller  607  will cause the guide link  604  and the piston  601  to also move laterally with respect to the longitudinal axis  602 . In this manner, the piston  601  may be aligned so as to move in cylinder  600  in a straight line that is substantially centered with respect to cylinder  600 . 
     FIG. 6 b  shows a side view of one embodiment of a precision alignment mechanism. A roller  607  is rotatably mounted on a locking eccentric  611  having a lower end  612  and an upper end  613 . The roller is mounted on a portion  610  of the locking eccentric  611  having a roller axis of rotation that is offset from the axis of rotation of the locking eccentric  611 . The lower end  612  is rotatably mounted in a lower bracket (not shown). The upper end  613  is rotatably mounted on an upper bracket  614 . FIG. 6 c  shows a perspective view of the embodiment shown in FIG. 6 b . The upper bracket  614  has a plurality of bracket holes  620  drilled through the upper bracket  614 . In a preferred embodiment, eighteen bracket holes are drilled through the upper bracket  614 . The bracket holes  620  are offset a distance from the axis of rotation of the locking eccentric  611  and are evenly spaced around the circumference defined by the offset distance. 
     FIG. 6 d  shows a top view of the embodiment shown in FIG. 6 b . The upper end  613  of the locking eccentric  611  has a plurality of locking holes  615 . The number of locking holes  615  should not be identical to the number of bracket holes  620 . In a preferred embodiment, the number of locking holes  615  is nineteen. The locking holes  615  are offset from the axis of rotation of the locking eccentric  611  by the same distance used to offset the bracket holes  620 . The locking holes  615  are evenly spaced around the circumference defined by the offset distance. FIG. 6 d  also shows a locking nut  616  that allows the locking eccentric  611  to rotate when the locking nut  616  is loose. When the locking nut  616  is tightened, the locking nut  616  makes a rigid connection between the locking eccentric  611  and the upper bracket  614 . FIG. 6 e  is the same view as shown in FIG. 6 d  but with the locking holes  615  shown. 
     During assembly, the piston is aligned in the following manner. The folded guide link is assembled with the locking nut  616  in a loosened state. The piston  601  (FIG. 6 a ) is aligned within the piston cylinder  600  (FIG. 6 a ) visually by rotating the locking eccentric  611 . As the locking eccentric  611  is rotated, the roller axis of rotation  608  (FIG. 6 a ) will be displaced both laterally and longitudinally to the guide link longitudinal axis  602  (FIG. 6 a ). The large lever ratio of the present invention requires only a very small displacement of the roller axis of rotation  608  (FIG. 6 a ) with respect to the longitudinal axis  602  (FIG. 6 a ) to align the piston  601  (FIG. 6 a ) within the piston cylinder  600  (FIG. 6 a ). In accordance with an embodiment of the invention, the maximum displacement range may be from 0.000 inches to 0.050 inches. In a preferred embodiment, the maximum displacement is between 0.010 inches and 0.030 inches. As the locking eccentric  611  is rotated, one of the locking holes  615  will align with a bracket hole  620 . FIG. 6 d  indicates such an alignment  630 . Once the piston  601  (FIG. 6 a ) is aligned in the piston cylinder  600  (FIG. 6 a ), a pin (not shown) is inserted through the aligned bracket hole and into the aligned locking hole thereby locking the locking eccentric  611 . The locking nut  616  is then tightened to rigidly connect the upper bracket  614  to the locking eccentric  611 . 
     In accordance with a preferred embodiment of the invention, a dual folded guide link piston linkage such as shown in cross-section in FIG.  7  and designated there generally by numeral  700  may be incorporated into a compact Stirling engine. Referring now to FIG. 7, pistons  701  and  711  are the displacer and compression pistons, respectively, of a Stirling cycle engine. As used in this description and the following claims, a displacer piston is either a piston without a seal or a piston with a seal (commonly known as an “expansion” piston). The Stirling cycle is based on two pistons executing reciprocating linear motion about 90° out of phase with one another. This phasing is achieved when the pistons are oriented at right angles and the respective connecting rods share a common pin of a crankshaft. Additional advantages of this orientation include reduction of vibration and noise. Additionally, the two pistons may advantageously lie in the same plane to eliminate shaking vibrations orthogonal to the plane of the pistons. While the invention is described generally with reference to the Stirling engine shown in FIG. 7, it is to be understood that many engines as well as refrigerators may similarly benefit from various embodiments and improvements which are subjects of the present invention. 
     The configuration of a Stirling engine shown in FIG. 7 in cross-section, and in perspective in FIG. 8, is referred to as an alpha configuration, characterized in that compression piston  711  and displacer piston  701  undergo linear motion within respective and distinct cylinders: compression piston  711  in compression cylinder  720  and displacer piston  701  in expansion cylinder  722 . Guide link  703  and guide link  713  are rigidly coupled to displacer piston  701  and compression piston  711  at piston connection points  702  and  712  respectively. Connecting rods  706  and  716  are rotationally coupled at connection points  705  and  715  of the distal ends of guide links  703  and  713  and to crankshaft  708  at crankshaft connection points  707  and  717 . Lateral loads on guide links  703  and  713  are substantially taken up by roller pairs  704  and  714 . As discussed above with respect to FIGS. 4 and 6, the pistons  701  and  711  may be aligned within the cylinders  720  and  722  respectively such using precision alignment of roller pairs  704  and  714 . 
     The devices and methods described herein may be applied in other applications besides the Stirling engine in terms of which the invention has been described. The described embodiments of the invention are intended to be merely exemplary and numerous variations and modifications will be apparent to those skilled in the art. All such variations and modifications are intended to be within the scope of the present invention as defined in the appended claims.