Piston-rod Device And Components Thereof

Abstract

In a piston for internal combustion engines having a circular head, two tapered part-way-round main skirts and two part-way-round vestigial skirts, a connecting rod with a clevised upper end is coupled to the piston by a wristpin mounted in a transverse bore formed in the central portion of a web structure extending between the main skirts to form an "H" configuration therewith. The four skirts are cloverleaf cam ground. The head has a single annular ring seating recess of which the bottom wall is of step configuration in radial cross section. An oil scraper ring is seated in the recess below and overhanging compression ring. Below the recess, the head circumference has two high-pressure oil drainage slots above the main skirts and two low-pressure drainage slots intervening and isolated from the high-pressure slots. The four slots are in communication with drainage basins formed radially inwards thereof in the bottom of the head by drain ports formed in the head to pass from such slots to such basin. The basins occupy four quadrants defined by the lateral web structure and by two ribs extending transversely from such structure to provide said vestigial skirts.

Description

This invention relates generally to pistons for internal combustion engines and to connecting rods, cylinders and other components associated with such pistons and to combinations of the elements just names. More particularly, this invention relates to a piston which is an improvement in respects over the piston disclosed in my copending application Ser. No. 669,905filed Sept. 22,1967, how abandoned but which also incorporates improvement features characterizing that earlier disclosed piston. Because of such relationship, that earlier filed copending application is hereby incorporated by reference to b e part of the disclosure hereof.

An object of this invention is to provide a piston characterized by one or more of the advantages of reduced mass, high strength and/or stiffness in relation to mass, facilitated cooling and efficient performance.

Another object of this invention is to provide in conjunction with a piston head a sealing assembly which occupies only a single circumferentially formed recess in the head but is highly effective in sealing against leakage past the head.

A further object of the invention is to provide new and improved skirt configurations for a piston characterized by part-way-round skirts.

A still further object of the invention is to provide improvements in the couplings of pistons and connecting rods and in such connecting rods themselves and in piston-cylinder combinations.

For a better understanding of how these and other objects of the invention are realized, reference is made to the following description of an exemplary embodiment of the invention and to the accompanying drawings wherein:

FIG. 1 is a full scale view in front elevation of a piston according to the invention, the piston being shown as received within a cylinder formed within an engine block and said piston being further shown coupled by a connecting rod to the crankshaft of an engine;

FIG. 2 is a view in side elevation of the FIG. 1 piston with the connecting rod being removed, the view of FIG. 2 being taken as indicated by the arrows 2-2 in FIG. 1;

FIG. 3 is a view in vertical cross section of the piston and associated components of FIG. 1, the FIG. 3 view being taken as indicated by the arrows 3-3 in FIG. 1;

FIG. 4 is another view in vertical cross section of only the piston of FIG. 1, the FIG. 4 view being taken as indicated by the arrows 4-4 in FIG. 1;

FIG. 5 is a bottom plan view of the FIG. 1 piston with the connecting rod removed;

FIG. 6 is a view, taken as indicated by the arrows 6-6 of FIG. 1, of a horizontal cross section taken through the head of the FIG. 1 piston and showing the cloverleaf cam grind of the piston skirts;

FIG. 7 is a view, taken as indicated by the arrows 7-7 of FIG. 1, of a horizontal cross section through the head of the FIG. 1 piston, the view showing details of the sealing assembly used with that head; and

FIGS. 8A--8D are enlarged schematic diagrams explanatory of the mode of operation of the sealing assembly of FIG. 7.

Referring now to the figures of which FIGS. 1--7 are to full scale, the reference numeral 10 designates (FIG. 1) a machined aluminum piston having a disc-shaped head 11 characterized by a circular cylindrical circumferential surface of which the radius (to the axis of the piston) is greater by a factor of at least three than the axial dimension of such surface. That is, head 11 is unusually small in the ratio of its axial thickness to its diameter.

Disposed on diametrically opposite sides of the head 11 are a pair of part-way-round main skirts 15 and 16 each having an angular extend around the head of less than 90.degree.. Each of the mentioned skirts is integral with and depends from a peripheral portion of the bottom of the head. The bottoms of skirts 15 and 16 are shown as being in contact with opposite sides of the wall of a cylinder 17 formed in an engine block 18, and in which the piston 10 is received.

Disposed below the bottom of head 11 and in integral relation with that bottom is a web means or web structure 20 extending diametrally and in the lateral direction between the skirts 15 and 16 to be integrally joined with the inner sides of such skirts. As best shown by FIG. 5, the web means 20 and the skirts 15 and 16 together form an "H" configuration wherein the end arms of the "H" are arcuate and provided by the skirts, and wherein the cross arm of the "H" is provided by the web means. As also shown by FIG. 5, the web means or structure 20 is symmetrical in relation to a diametral plane 19 passing through the piston axis and through the skirts 15,16 to serve as a lateral center plane for structure 20. Structure 20 is a unitary structure in the sense that it is continuously solid in the transverse direction (i.e., in the direction normal to the lateral extent of the structure) from one to the other of the transversely opposite outwardly presented faces of the structure.

As illustrated by FIGS. 1, 3 and 5, the web structure 20 has a laterally central portion 21 which is transversely thickened (in a symmetrical manner about the mentioned center plane 19) relative to the transverse thickness of the portions of such structure which lie on opposite lateral sides of the central portion. The lower part of such portion 21 is in the form of a downstanding block or stanchion 22 having parallel laterally extending vertical faces 23 and 24 (FIGS. 3 and 5) on transversely opposite sides of the block. Formed within the block 22 is a transverse cylindrical bore 25 extending horizontally from face 23 to face 24 at a level such that there is substantial room or clearance between the top of the bore and the portions of the bottom of head 11 which overhang the ends of the bore. The apertured block or stanchion 22 serves as a wristpin boss in that a wristpin 26 is received in bore 25 to have opposite ends of the pin project outwardly in transversely opposite directions from the walls 23 and 24 of the boss.

To either side of portion 22, the web structure 20 narrows down in transverse thickness to assume the form of thin plate portions at their laterally outer ends are joined to the skirts 15 and 16 by transition portions 30 and 31 characterized by transversely opposite fairing surfaces 32, 33 and 34, 35 which are part-circular-cylindrical surfaces, and which curve away from each other with distance in the radially outward direction to thereby progressively increase in such direction the transverse cross section of each of those transition portions. The transition portions 30 and 31 serve (a) to stiffen central regions of skirts 15 and 16 which are of relatively wide angular extent and (b) to transmit sidewise or lateral load from those wide regions to the relatively narrow plate portions 28 and 29. At the same time, the thinness of the plate regions 28 and 29 serve to lighten the mass of the piston.

Returning to the steel wristpin 26, the transversely opposite ends of that pin are received in separate corresponding cylindrical bores 37 and 38 formed in a pair of fork arms 39 and 40 (FIG. 3) on transversely opposite sides of block 22 and constituting parts of the clevised top end of an aluminum connecting rod 45. The pin 26 thus serves to couple piston 10 to such rod. Because the bore 25 in stanchion 22 is disposed well below the bottom portions of head 11 which overhang the ends of that bore, adequate room is provided beneath the bottom of the head for the upper ends of the fork arms 39 and 40 to encircle the pin 26.

To prevent accidental uncoupling of the piston and rod by horizontal sliding of the pin 26 out of the bores in which it is received, the pin has at one end a flange 46 of larger diameter than such bores and, at the other end, an annular slot 47 in which is received a selectively removable split retaining ring 48 of larger diameter than such bores.

As illustrated in FIG. 3, the upper parts of the fork arms 39 and 40 of rod 45 are set transversely outward from the shank 51 of that rod and are coupled to such shank by faired transition portions 52 and 53 forming a "Y" configuration with the shank. The described transverse spreading of the fork arms has the advantage that stanchion 22 may be of the calculated transverse cross section dimension necessary to withstand the stresses generated by loads transmitted between the piston and the shank 51 while, concurrently, it is not necessary to accommodate such transverse dimension for stanchion 22 by increasing the transverse cross-sectional dimension of shank 51 to a value equal to the transverse distance between the outer faces of the upper parts of fork arms 39 and 40. That is, it is not necessary to increase the transverse dimension of shank 51 to a value greater than that needed to provide a cross section for the shank which enables the shank to withstand the mentioned stresses.

Apart from the features just discussed of rod 45, that rod is similar to the rod disclosed in my copending application Ser. No. 709,980 which was filed Mar. 4,1968, how U.S. Pat. No. 3,482,468,and of which the disclosure is hereby incorporated by reference to be made a part of the present disclosure. Among the noteworthy aspects of such rod as disclosed in copending application Ser. No. 709,980 are the features that the shank portion of the rod is of "H" configuration in cross section and that the part of the shank which provides the cross arm of the "H" is hollow tubular.

The shank 51 of rod 45 is coupled by a split journal 55 and fastening screws 56 for the journal to a crank arm rod bearing 57 on a crankshaft 58 of which the main bearings 59 are immersed in a body of oil 60 contained within a crankcase (not shown). The mode by which rod 45 converts the reciprocating motion of piston 10 within the cylinder 17 into rotary motion of the crankshaft 58 is so well known that it need not be discussed in detail herein.

To improve the stability of the piston in the course of such reciprocating motion, the piston 10 is provided with part-way-round vestigial skirts 65 and 66 disposed on diametrally opposite sides of head 11 circumferentially between the main skirts 15 and 16. Like elements 15 and 16 the vestigial skirts each have an angular extent around head 11 of less than 90.degree.but the vestigial skirts are very much smaller in axial extent than are the main skirts.

The skirts 65 and 66 provided by the radially outward ends of a pair of downwardly salient ribs 67 and 68 (FIG. 5) formed on the bottom of head 11 and extending transversely in opposite directions from the center portion 21 of web means 20 to the peripheral region of the head. With distance radially outward, the ribs 67 and 68 first contract in lateral cross section down to respective neck portions 69 and 70. Beyond such neck portions, the mentioned ribs have respective faired portions 71 and 72 which progressively increase in lateral cross section in the direction toward the circumference of head 11. The laterally opposite sides of such faired portions are in the form of fairing surfaces 73,74 and 75,76 which curve away from each other with progress in the last-named direction. The bottom surfaces 77 (FIGS. 1 and 3) of ribs 65 and 66 are directly above the fork arms 39,40 of rod 45 but are spaced far enough above bore 25 and the pin 26 therein to make room for such fork arms to encircle that pin.

The "X" configuration made by ribs 67,68 with the web structure 20 serves to stiffen the head 11 in both of its orthogonal dimensions. Further, the "hour glass" shape (in horizontal cross section) of each of ribs 67,68 serves to provide adequate rearward stiffening for vestigial skirts 65,66 while concurrently making the piston head 11 lighter than if such ribs were to be of straight line configuration. Still further, that "hour glass" shape implements (as later described) the drainage from the piston of oil which the motion of crankshaft 58 has splashed up from reservoir 60 onto the cylinder wall 17 and on to the underside of the piston for the purposes of, respectively, lubricating the piston and cooling the piston.

Sealing is provided between the piston and the cylinder wall by ring means 80 received within a single annular recess 81 formed in the circumference of head 11. The particular structure of the sealing combination provided by ring means 80 and the recess 81 will be later described in detail. For the present, it is merely pointed out that a factor contributing to the success of the "single zone" sealing means 80,81 is that piston 10 provides for full and easy drainage back to reservoir 60 of oil accumulating beneath the ring means 80. The features of the piston which implement such drainage are as follows.

During each revolution of the crankshaft 58, the resulting horizontal displacements of the bottom of rod 45 are productive of tilting of that rod to cause first one and then the other of skirts 15 and 16 to be forced sidewise against cylinder wall 17 to thereby subject to high pressure the oil disposed between such wall and said skirts. To provide for drainage of that high pressure oil, head 11 has formed in the circumference thereof a pair of high pressure drainage slots 85 and 86 disposed above skirts 15 and 16, respectively, such that each of those slots is below the recess 81. Furthermore, the head 11 has formed in the circumference thereof a pair of low-pressure oil drainage slots 87 and 88 disposed on diametrically opposite sides of the head to be circumferentially between the high-pressure slots. As best shown in FIG. 1, the high-pressure and low-pressure slots are isolated from each other endwise by unrecessed circumferential portions 89 of the head 11. The lower wall of the low-pressure slots 87,88 is provided by an edge flange 84 which acts as a scraper for low-pressure oil, and of which the bottom side marks the bottom proper of the head 11.

Oil in the slots 85--88 is drained from the exterior of the head through a plurality of drain ports 90 spaced angularly around the head and passing from the slots through the metal of the head to the appropriate one of four drainage basins 92--95 (FIG. 5), disposed radially inward of the slots. Those four drainage basins are each in the form of a shallow concave recess formed in the bottom of head 11 in a respective one of the four quadrants into which the bottom of the head is divided by the "X" configuration defined by the web structure 20 and the ribs 67,68. As illustrated by FIG. 5, the described fairing surfaces of the faired portions 30,31 of structure 20 and of the faired portions 71 and 72 of the ribs also form side surfaces for the mentioned drainage basins to thereby enable ones of the drain ports 90 to pass through such faired portions and yet open onto an appropriate basin. Thus, the employment of the described faired portions provides the additional advantage of enlarging the drainage basins (to result in better drainage and better cooling of the head from the underside by splashed oil) and of enabling drain ports to be angularly placed closer together in the circumferential regions of head 11 which are outwards of, respectively, the web structure 20 and the ribs 67,68 to thereby effect better oil drainage in such regions.

Considering now some further details of the described skirts, the main skirts 15 and 16 have outer faces 105 and 106 which are periodically interrupted by axially spaced horizontal antigalling grooves 107, but which are considered for purposes of further discussion as being continuous. Those skirt faces 105 and 106 have angularly central zones 108 and 109 (FIG. 6) conforming to a surface of revolution 110 defined around the axis of the piston. In the course of forming the piston by machining, portions of the original faces 105,106 to either angular side of the central zones 108,109 are ground away. As a result, the skirt faces 105,106 in the finished piston are each characterized on angularly opposite sides of the central zones 108,109 of such faces by face expanses 111,112 and 113,114 which are spaced radially inwards of the geometric surface 110. To give an idea of how much the side portions of the faces 105 and 106 are ground down, in the represented piston, the diametral distance between the tip of the upper face portion 111 of skirt 15 and the tip of the lower face portion 114 of skirt 16 is less in value by 0.002 inch than the diametral distance between the central face portions 108 and 109 of those two skirts.

In like manner, the vestigial skirts 65 and 66 have outer faces 115,116 characterized by angularly central zones 117,118 which conform to surface 110 and by expanses 119,120 and 121,122 which are disposed on angularly opposite sides of those central zones and are spaced radially inwards of the geometric surface 110.

The shaping, as described, of the outer faces of the four skirts is known in the art as a "cloverleaf cam grind" because such outer "cam" faces conform to and form parts of a closed four-lobed curve 125 which, when having pronounced lobes, resembles a cloverleaf In the prior art, a cloverleaf cam grind has been imparted to the exterior of a full-way-round or 360.degree. piston skirt in order to provide room into which parts of such skirt can expand as the diametral size of the skirt tends to increase or "grow" by heating of the piston, but the skirt is concurrently inhibited from growing by contact of the skirt with the restraining wall of the cylinder in which the piston is received. In the presently disclosed piston, the discontinuities around the head between the part-way-round skirts serve of themselves to provide room for expansion of the skirts in response to heating of the piston. As opposed, however, to making room for skirt expansion, the cloverleaf cam grind performs in the present piston a function which it does not perform in a piston with a full-way-round skirt, namely, that of assuring, when the piston is full "grown," a clearance between the side portions of the skirts and the cylinder wall to thereby permit an adequate amount of lubricating oil to be maintained between the skirts and that wall. Moreover, the cloverleaf cam grind minimizes the areas of contact between the piston and cylinder wall to thereby reduce friction.

The four skirts 15,16 and 65,66 preferably characterized by such a cloverleaf cam grind throughout the axial length of each.

As another feature of main skirts 15 and 16, the mentioned geometric surface of revolution 110 is preferably not a circular cylindrical surface but, instead, is a frustoconical surface which progressively diminishes in cross section in the direction from the bottoms to the tops of skirts 15,16 and over the axial length of such skirts. Accordingly, the central outer face zones 108 and 109 of the main skirts convergently taper towards each other in that direction and over the skirt lengths. The amount of taper is such that the diametral distance between the central outer face portions 108, 109 at the bottoms of skirts 15,16 is 0.005 inch greater than at the tops of such skirts. Moreover, the diametral distance between such outer face portions at the bottoms of skirts 15,16 is greater by 0.0005 inch than the inner diameter of cylinder 17 when the piston and cylinder are at room temperature, and when the piston is outside the cylinder to be unstressed by contact with the wall thereof. That is, piston 10 at the bottom of skirts 15,16 has a minus clearance when fitted at room temperature into the cylinder 17.

The described tapering of the main skirts of the piston compensates for the greater lateral expansion of the top of the piston than the bottom thereof in response to heating of the piston from the top by hot gases in the combustion chamber above the piston. Moreover, the minus clearance at room temperature of the piston at the bottom of the skirts thereof (and in relation to cylinder 17) promotes stability of alignment of the piston in the cylinder when the engine is cold. As a practical matter, the ability to so taper the main skirts and to provide for a minus clearance at room temperature is dependent in large part upon the lateral stiffening provided for the skirts 15,16 by the web structure 20 and upon the consequent accurate predictability of how much the piston will laterally grow from point to point over its axial length (and within cylinder 17) when the engine becomes hot.

Turning now to the earlier mentioned sealing combination 80,81, such combination is best shown by FIGS. 7 and 8A--8D. In that combination, the annular recess 81 is comprised of an upper annular groove 130 extending radially into the head 11 from the circumference thereof and having an upper sidewall 131 and a lower sidewall 132. At its radially outward opening, recess 81 is enlarged to an axial width greater than that of the radially inward end of groove 130 by an annular channel 135 extending axially downward from wall 132 and extending radially into head 11 from the circumference thereof. As illustrated, channel 135 serves to shorten the radial extent of wall 132 relative to that of wall 131. Moreover, channel 135 also renders the bottom of recess 81 of step configuration in radial cross section.

Disposed in recess 81 is a resilient steel compression ring 140 having a radially inward portion seated in groove 130 between walls 131,132, the said ring also having a radially outward portion overhanging the channel 135. Ring 140 is slightly less in axial width than the groove 130, wherefore such ring is received into the groove with a slight clearance.

The ring 140 is a split resilient ring having opposite peripheral ends separated by a gap 141 (FIG. 7). The ring is eccentrically moveable in relation to the piston axis and is of greater outer diameter than the piston head either when the head is outside cylinder 17 (so that the ring is relaxed and gap 141 has maximum size) or when the head is received within the cylinder. In the latter instance, ring 140 is resiliently compressed by the cylinder wall so that the gap 141 is smaller in size than when the ring is relaxed. When head 11 is pressed against one lateral side of the cylinder wall as a result of, say, a lateral component of force exerted by rod 45 (when in tilted position) against the head, the portion of ring 140 on that side of the head is forced deeper in the groove 130, and the ring as a whole becomes eccentric i(or more eccentric) in relation to the axis of the head. Groove 130 is deep enough to permit such portion of the ring to be forced fully into the recess 81. Disposed radially inward of ring 140 in groove 130 is a polygonal resilient steel ring-shaped leaf spring 145 extending circumferentially around the inside of the groove. When ring 140 is in centered relation with head 11, spring 145 contacts both the inside of the ring and the inner wall 146 of groove 130 but is concurrently in a slack or relaxed condition. When, however, a particular circumferential portion of ring 140 is forced deeper into groove 130 as described, then the section of spring 145 to the inside of such portion becomes compressed to urge that portion outward. In this way, the leaf spring 145 tends to maintain ring 140 in centered relation with the head and, thus, to maintain the head in centered relation with the axis of cylinder 17.

Received in channel 135 beneath the overhanging portion of ring 140 is a resilient steel oil scraper ring 150. Like ring 140, the scraper ring 150 is a split ring having opposite peripheral ends separated by a gap 151 (FIG. 7) which is angularly displaced from the gap 141 in ring 140 to close any clear passage for oil through the peripheral region of the composite ring structure. Moreover, like ring 140 the scraper ring is of greater outer diameter than head 11, is resiliently compressed to narrow the size of its gap 151 when head 11 is received within cylinder 17, is eccentrically moveable in relation to head 11, and is adapted to have a side portion thereof forced fully into channel 135 when the head 11 is pressed hard on that side against the cylinder wall. The scraper ring 150 is, however, of smaller radial and axial width than the compression ring 140. Further, the scraper ring preferably has the same axial width as the channel 135. Hence, when the compression ring 140 is bearing under pressure against the lower sidewall 132 of groove 130, substantially no clearance exists between the ring 140 and the side enclosure for such ring which is formed by the overhanging portion of ring 140 and the lower sidewall 155 of channel 135.

As already indicated, the scraper ring 150 has an axial width not exceeding that of channel 135 but preferably equal to that of channel 135. While the ring 150 may have an axial width slightly less than that of channel 135, the difference (clearance) between the respective axial widths of ring 140 and groove 130 is always greater than the difference (clearance) between the respective axial widths of ring 150 and channel 135.

FIGS. 8A--8D illustrate the operation of the described sealing combination during all of the various strokes of the piston.

In FIG. 8A, the piston is being driven down in the course of a power stroke. The pressure of the combustion gases (in the chamber above the piston) act upon the compression ring 140 to maintain it in pressure contact with the lower wall 132 of groove 130 despite the tendency of inertia forces to lift ring 140 off that wall. In opposition, however, to such pressure from the combustion gases, the downward stroke of the piston and any slapping of the piston against the cylinder wall of a high hydraulic pressure tending (along with the inertia forces) to lift ring 140 off the wall 132 and to force the oil through the interface 160 between such ring and that wall. In the situation shown in FIG. 8A, however, ring 150 is driven against ring 140 by inertia forces to preclude any such flow of oil between those two rings. Also, since there is substantially no clearance between the lower side of ring 150 and the wall 155 of channel 135, little or no oil can flow through the interface 165 between ring 150 and wall 155. Hence, during the power stroke, the amount of oil or gas which leaks past the inside of the rings 140 and 150 is minimal.

In the upper exhaust stroke represented by FIG. 8B, inertia forces and the pressure of the gases being exhausted from the combustion chamber cause a driving downward of both ring 140 and ring 150 relative to the head 11. Therefore, the oil and/or gas leakage past the inside of the rings is minimal for the reasons explained in connection with FIG. 8A.

During the downward intake stroke shown by FIG. 8C, both of rings 140 and 150 are driven upward relative to head 11 by inertia forces acting on the rings and by vacuum forces in the combustion chamber. Again, there is generated in the film of oil below the rings, a high hydraulic pressure tending to force oil through the now open interface 160. In order, however, to reach such interface, the oil must first flow through the narrow labyrinthine path initially defined through interface 165 (between the lower side of ring 150 and lower channel wall 155) and then through the interface 166 defined between the radially inward wall of ring 150 and the radially inward wall 167 of channel 135. Since that labyrinthine path offers high resistance to oil flow, and since any oil which flows through that path and then through interface 160 must pass through the interface 168 between wall 131 and ring 140 in order to leak past the whole ring structure, very little oil flows past the insides of the rings during the intake stroke.

In the upper compression stroke depicted by FIG. 8D, the rings 140 and 150 are caused by inertia forces and the pressure of the gas being compressed in the combustion chamber to occupy the same positions relative to head 11 as they do in FIGS. 8A and 8B. In the course, therefore, of the compression stroke, a minimal amount of oil flows past the insides of the rings.

From what has been said, it will be seen that ring 150 acts in relation to ring 140 as a "guard ring" which operates during the power, exhaust and compression strokes to prevent oil from flowing through the interface 160 and which, further, operates during the intake stroke to interpose a high resistance path to the flow of such oil. The use of a sealing combination of the sort described has been found efficient to seal against oil leakage to the point where the engine can be run upside down (i.e., with its crankcase above the piston) without smoking.

Some other reasons why the described sealing combination provides an excellent sealing action are as follows.

First, the clearance required between a conventional piston ring and the groove in which such ring is seated is, in general, proportional to the radial depth occupied by the ring in the groove. In the combination shown by FIGS. 7 and 8A--8D, however, the lower sidewall 132 of groove 130 is rendered much shorter (by the presence of channel 135) than the upper sidewall 131 thereof. Hence, the clearance required between ring 140 and the groove walls 131 and 132 is substantially less than if wall 132 were extended out to the circumference of head 11. Further, no clearance tolerance need be provided in connection with the fitting of scraper ring in channel 135 between the lower channel wall 155 and the overhanging portion of ring 140. Accordingly, the described sealing structure permits rings 140 and 150 to be seated in recess 81 with a clearance between the sidewalls of the recess and the received ring structure (consisting of the two rings 140 and 150) which is much reduced (e.g., by about 50 percent) relative to the clearance which would normally be required for the maximum radial depth which the ring structure occupies in that recess. Manifestly, however, such reduced clearance reduces the leakage of oil and/or gases past the inside of the ring structure.

Second, during the power, exhaust and compression strokes, the pressure of gas in the combustion chamber may tend to bind ring 140 in an eccentric position on its seat 132 such that, on one lateral side of the head, the corresponding side portion of ring 140 is received fully in groove 130. In those circumstances and in the absence of scraper ring 150, there would be no effectual sealing element of such lateral side, and oil disposed on that side between the piston and cylinder wall would be forced upward past the head by the high hydraulic pressure generated on that oil when the head is urged on that side towards the wall by the lateral component of force created by the inclination from vertical of the connecting rod 45 as its bottom end follows the rotary motion of the crankarm bearing 57 (FIG. 1). In the described sealing combination, however, the compression ring 140 is interposed between the gas pressure and the scraper ring 150 to protect the scraper ring from being bound in position relative to head 11 by the said gas pressure. Therefore, ring 150 is always enabled to provide a sealing action preventing or much reducing the leakage of oil and/or gas past the outside of the described sealing combination.

For the shown piston which has a head diameter of 2 9/16 inches, typical dimensional values for the considered sealing structure are as follows:

EXAMPLE __________________________________________________________________________ Ring 140 Radial width 0.0100" Axial thickness 0.0075" Ring 150 Radial width 0.0075" Axial thickness 0.0024" Spring 145 Radial thickness 0.0015" Axial width 0.0075" Groove 130 Radial depth 0.0145" Axial width 0.0076" Channel 135 Radial depth 0.0080" Axial width 0.0024" __________________________________________________________________________

For pistons having larger or smaller head diameters than the 2 9/16 inches diameter to which the values of the foregoing example apply, such values may be scaled up or down in proportion to the head diameter. Moreover, while the dimensional values which have been given for a piston of 2 9/16 inches head diameter yields highly satisfactory results, such values are exemplary only and may be departed from to a considerable degree while being characteristic of a sealing structure in accordance with the invention.

Some yet unmentioned advantages of the described piston are as follows.

It has been proposed in the prior art that pistons with part-way-round skirts should have skirts which are flexible in order to reduce engine noise and avoid the necessity of close piston-to-cylinder fits. A disadvantage, however, of such proposal as applied to aluminum pistons is that aluminum is a brittle metal which tends to break if flexed too much or for too long. Accordingly, the main skirts of the presently described piston are rendered very stiff by the web structure 20 interposed between them. Such structure is particularly apt for imparting skirt rigidity because, due to the cloverleaf cam grind of the skirts, lateral forces exerted by the cylinder wall and acting on the piston are concentrated in the narrow central zones 108,109 of the skirts and, hence, are directly in line with the opposing reaction forces exerted on the piston skirts from the web structure. There is, therefore, little or no angular flexing of the skirts in horizontal cross-sectional planes. Moreover, since the active and reactive piston forces caused by lateral loading of the piston are in normal relation to the wristpin, there is no interaction between such lateral loading forces and the forces involved in the transmission of load through the connecting rod between the piston and the crankshaft.

The described taper and minus clearance of the skirts promotes stability of the piston in the cylinder when the engine is cold. Also, such taper provides that, as the engine warms up and the piston is heated by the hot gases in the combustion chamber, the differential growth induced over the axial length of the piston by the temperature differential in the piston over such length will produce an optimum fit (but not too close a fit at any point) between the piston and the cylinder when the engine is hot. For such hot condition, the central zones 108,109 of the skirts approach close to conforming to a circular cylindrical surface of revolution about the piston axis. Whether, however, the engine is cold or hot, the cloverleaf cam grind of the vestigial and main skirts ensures low friction between the cylinder and piston with accompanying good lubrication of the piston.

Because of the effective sealing action provided by the described composite ring means and the step configuration recess and because of the features of the piston which permit very free drainage of oil back into the crankcase (to thereby prevent accumulation and pressurizing of oil beneath the ring sealing means and a consequent forcing of the oil under high hydraulic pressure past such sealing means), the piston head 11 need only have one ring-seating recess. Accordingly, the head avoids the disadvantages of prior art pistons having piston rings in a plurality of axially spaced recesses, such disadvantages being that oil is trapped between the spaced rings to become overheated and to be forced by high hydraulic pressure past the upper ring when the piston rocks or "cocks" in the cylinder or slaps against the cylinder wall. Since, for the reasons discussed heretofore, the head 11 need only have one ring-seating recess, the head can be of relative small axial thickness (i.e. of an axial thickness which is only between 3 and 4 times the axial width of the recess itself) to thereby be of low mass. In turn, that low mass reduces the inertia of the piston and implements the cooling thereof by oil splashed up from the crankcase onto the underside of the head. In this latter connection, the single web structure 20 and the radially central cross sections of reduced thickness of such structure and of ribs 65,66 serve to provide maximum exposure of the hottest part of the bottom of the head so as to further implement the cooling of the head by upwardly splashed oil. Further, in the latter connection, the described basins 92--95 not only promote oil drainage but also reduce the mass of the head and increase (due to their concave shape) the head' bottom surface area exposed to cooling oil while, concurrently, the "arch" shape in vertical cross section of the basins serves to maintain the strength and stiffness of the head despite the loss therefrom of the material removed in the course of forming the basins. Moreover, despite the thinness of head 11 in the axial direction, the reinforcing "X" configuration formed by web structure 20 and ribs 65,66 serves to provide additional stiffness and strength for the thin head in the presence of high pressures and loads acting thereon.

In connection with the foregoing, it will be understood that the present invention hereof is applicable to pistons other than those made of aluminum and to pistons other than those designed for use in internal combustion engines. Accordingly, the invention set forth herein is not to be considered as limited save as is consonant with the recitals of the following claims.

Claims

I claim:

1. Apparatus comprising, a piston head having formed in the circumference thereof a single annular recess, said recess being comprised of an annular groove having upper and lower sidewalls extending radially into said head from the circumference thereof, and said recess being further comprised of an annular channel enlarging the radially outward opening of said recess relative to the axial width of said recess at the portion overhanging portion. radially inner end of said groove, said channel extending axially downward from said lower wall and radially into said head from the circumference thereof to a lesser depth than the radially inward extend of said groove to thereby shorten the radial extent of said lower groove sidewall relative to that of said upper groove sidewall and to render the bottom of said recess of step configuration in radial cross section a single first split resilient sealing ring having a radially inward portion received with axial clearance between said upper and lower sidewalls of said groove, and having a radially overhanging said channel, and a single split resilient sealing ring received in said channel beneath said overhanging portion of said first ring and in abutting engagement with said overhanging portion.

2. The improvement as in claim 1 in which said first ring has a greater inner diameter than said groove to be eccentrically movable therein, and in which said second ring has a greater inner diameter than said channel to be eccentrically movable in said channel.

3. The improvement as in claim 2 further comprising resilient spring means extending circumferentially around said groove radially inwards of said first ring, said spring means being adapted to urge said first ring into centered relation with the axis of said head.

4. The improvement as in claim 3, wherein the resilient spring means includes an elongated resilient polygonal leaf spring inserted in said recess, said leaf spring being relaxed but in touching contact with both said ring and the inner wall of said recess when said ring is relaxed.

5. In the combination comprising a metal piston having a circular head and a pair of part-way-round main skirts disposed on diametrically opposite sides of said head and integral with and depending from respective peripheral portions of the bottom of said head, the improvement comprising, web means integral with and depending from the bottom of said head and extending diametrally beneath said bottom between the inner sides of said skirts to be integrally joined with said skirts, said web means forming with said skirts an "H" configuration wherein the end arms of said "H" are arcuate and provided by said skirts and the cross arm of said "H" is provided by said web means; a single annular ring-seating recess formed in the circumference of said head, said recess including an annular groove having upper and lower sidewalls and extending radially into said head from the circumference thereof, and an annular channel axially enlarging the radially outward opening of said recess relative to the axial width of said recess at the radially inner end of said groove, said channel extending axially downward from said lower wall and radially into said head from the circumference thereof to a lesser depth then the radially inward extent of said groove, and said channel thereby rendering the bottom of said recess of step configuration in radial cross section; a split, resilient compression ring having a radially inward portion received between said sidewalls of said groove and having a radially outward portion overhanging said channel, the axial width of said compression ring being less than the axial width of said groove; and a split resilient oil scraper ring received in said channel beneath said overhanging portion of said compression ring, the axial width of said scraper ring being at most equal to that of said channel, and the difference between the respective axial widths of said compression ring and groove being greater than the difference, if any, between the respective axial widths of said scraper ring and channel.

6. Apparatus comprising, a piston head having formed in the circumference thereof a single annular recess, said recess being comprised of an annular groove having upper and lower sidewalls extending radially into said head from the circumference thereof, and said recess being further comprised of an annular channel enlarging the radially outward opening of said recess relative to the axial width of said recess at the radially inner end of said groove, said channel extending axially downward from said lower wall and radially into said head from the circumference thereof to a lesser depth than the radially inward extend of said groove to thereby shorten the radial extent of said lower groove sidewall relative to that of said upper groove sidewall and to render the bottom of said recess of step configuration in radial cross section; a first split resilient sealing ring having a radially inward portion received with axial clearance between said upper and lower sidewalls of said groove, whereby the axial width of said first ring is less than that of said groove, and having a radially outward portion overhanging said channel; and a second split resilient sealing ring received in said channel beneath said overhanging portion of said first ring, the axial width of said second ring being at most equal to that of said channel, and the difference between the respective axial widths of said first ring and groove being greater than the difference, if any, between the respective axial widths of said second ring and channel.

7. In the combination comprising a metal piston having a circular head and a pair of part-way-round main skirts disposed on diametrically opposite sides of said head and integral with and depending from respective peripheral portions of the bottom of said head, the improvement comprising web means integral with and depending from the bottom of said head and extending diametrically beneath said bottom between the inner side of said skirts to be integrally joined with said skirts, said web means forming with said skirts an "H" configuration wherein the end arms of said "H" are arcuate and provided by said skirts and the cross arm of said "H" is provided by said web means, a single annular ring-seating recess formed in the circumference of said head, said recess including an annular groove having upper and lower sidewalls and extending radially into said head from the circumference thereof, and an annular channel axially enlarging the radially outward opening of said recess relative to the axial width of said recess at the radially inner end of said groove, said channel extending axially downward from said lower wall and radially into said head from the circumference thereof to a lesser depth than the radially inward extent of said groove, said channel thereby rendering the bottom of said recess of step configuration in radial cross section, a single split resilient compression ring having a radially inward portion received between said sidewalls of said groove and having a radially outward portion overhanging said channel, and a single split resilient oil scraper ring received in said channel beneath said overhanging portion of said compression ring and in abutting engagement with said overhanging portion.

8. In combination with a piston for an internal combustion engine having a circular head and a pair of part-way-round main skirts disposed on diametrically opposite sides of said head and integral with and depending from respective peripheral portions of the bottom of said head, said piston being adapted for operation thereof with a split resilient compression ring and a split resilient oil scraper ring, improved sealing means comprising one and only one split resilient compression ring, one and only one annular ring-seating recess formed in the circumference of said head, said recess including an annular groove having upper and lower sidewalls and extending radially into said head from the circumference thereof, said annular groove receiving a radially inward portion of said split resilient compression ring between said sidewalls and having an axial width greater than the axial width of said compression ring, and one and only one split resilient oil scraper ring, said recess having an annular channel axially enlarging the radially outward opening of said recess relative to the axial width of said recess at the radially inner end of said groove, said channel extending axially downward from said lower wall and radially into said head from the circumference thereof to a lesser depth than the radially inward extent of said groove, thereby rendering the bottom of said recess of step configuration in radial cross section, said annular channel receiving said split resilient oil scraper ring and having an axial width at least equal to that of said oil scraper ring and said compression ring having a radially outward portion overhanging said channel, and said split resilient oil scraper ring being received in said channel beneath said overhanging portion of said compression ring.

9. The improvement according to claim 8, wherein the difference between the respective axial widths of said compression ring and groove is greater than the difference, if any, between the respective axial widths of said scraper ring and channel.