US3111261A - Rotor and bearing construction for rotary mechanisms - Google Patents

Rotor and bearing construction for rotary mechanisms Download PDF

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Publication number
US3111261A
US3111261A US26233A US2623360A US3111261A US 3111261 A US3111261 A US 3111261A US 26233 A US26233 A US 26233A US 2623360 A US2623360 A US 2623360A US 3111261 A US3111261 A US 3111261A
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Prior art keywords
rotor
sleeve
bearing
clearance
eccentric
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US26233A
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English (en)
Inventor
Bentele Max
Jones Charles
Alexander H Raye
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Curtiss Wright Corp
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Curtiss Wright Corp
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Priority to US26233A priority Critical patent/US3111261A/en
Priority to DEC23918A priority patent/DE1158750B/de
Priority to CH472161A priority patent/CH383070A/de
Priority to BE603137A priority patent/BE603137A/fr
Priority to GB15683/61A priority patent/GB939704A/en
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Publication of US3111261A publication Critical patent/US3111261A/en
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    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16C—SHAFTS; FLEXIBLE SHAFTS; ELEMENTS OR CRANKSHAFT MECHANISMS; ROTARY BODIES OTHER THAN GEARING ELEMENTS; BEARINGS
    • F16C17/00—Sliding-contact bearings for exclusively rotary movement
    • F16C17/12—Sliding-contact bearings for exclusively rotary movement characterised by features not related to the direction of the load
    • F16C17/22—Sliding-contact bearings for exclusively rotary movement characterised by features not related to the direction of the load with arrangements compensating for thermal expansion
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F02—COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
    • F02B—INTERNAL-COMBUSTION PISTON ENGINES; COMBUSTION ENGINES IN GENERAL
    • F02B53/00—Internal-combustion aspects of rotary-piston or oscillating-piston engines
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F02—COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
    • F02B—INTERNAL-COMBUSTION PISTON ENGINES; COMBUSTION ENGINES IN GENERAL
    • F02B53/00—Internal-combustion aspects of rotary-piston or oscillating-piston engines
    • F02B2053/005—Wankel engines
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F02—COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
    • F02B—INTERNAL-COMBUSTION PISTON ENGINES; COMBUSTION ENGINES IN GENERAL
    • F02B2730/00—Internal-combustion engines with pistons rotating or oscillating with relation to the housing
    • F02B2730/01—Internal-combustion engines with pistons rotating or oscillating with relation to the housing with one or more pistons in the form of a disk or rotor rotating with relation to the housing; with annular working chamber
    • F02B2730/018—Internal-combustion engines with pistons rotating or oscillating with relation to the housing with one or more pistons in the form of a disk or rotor rotating with relation to the housing; with annular working chamber with piston rotating around an axis passing through the gravity centre, this piston or the housing rotating at the same time around an axis parallel to the first axis
    • Y—GENERAL TAGGING OF NEW TECHNOLOGICAL DEVELOPMENTS; GENERAL TAGGING OF CROSS-SECTIONAL TECHNOLOGIES SPANNING OVER SEVERAL SECTIONS OF THE IPC; TECHNICAL SUBJECTS COVERED BY FORMER USPC CROSS-REFERENCE ART COLLECTIONS [XRACs] AND DIGESTS
    • Y02—TECHNOLOGIES OR APPLICATIONS FOR MITIGATION OR ADAPTATION AGAINST CLIMATE CHANGE
    • Y02T—CLIMATE CHANGE MITIGATION TECHNOLOGIES RELATED TO TRANSPORTATION
    • Y02T10/00—Road transport of goods or passengers
    • Y02T10/10—Internal combustion engine [ICE] based vehicles
    • Y02T10/12—Improving ICE efficiencies

Definitions

  • the present invention relates to a rotor construction for rotary mechanisms using a lightweight metal alloy or other material which preferably has good heat conducting properties, such as a magnesium or aluminum alloy; and more particularly to means which make the use of such a rotor feasible, with consequent savings in weight and advantages in cooling efiiciency.
  • the present invention is particularly useful in rotating combustion engines of the type which comprise an outer body having an axis, axially-spaced end walls, and a peripheral wall interconnecting the end walls; the inner surfaces of the peripheral wall and end walls form a cavity, and an inner body or rotor is mounted within the cavity between its end walls.
  • the inner surface of the peripheral wall is preferably parallel to the axis of the cavity and, as viewed in a plane transverse to this axis, the inner surface preferably has a multi-lobed profile which is substantially an epitrochoid.
  • the axis of the rotor is eccentric from and parallel to the axis of the cavity of the outer body, and the rotor has axially-spaced end faces disposed adjacent to the end walls of the outer body, and a plurality of circumferentiallyspaced apex portions.
  • the rotor is rotatable relative to the outer body such that the apex portions substantially continuously engage the inner surface of the outer body to form a plurality of working chambers which vary in volume during engine operation, as a result of relative rotation between the rotor and the outer body.
  • Such engines also include an intake passage for administering a fuel-air mixtu e to the chambers, an exhaust port for the chambers, and suitable ignition means so that during engine operation the worxing chambers of the engine undergo a cycle of operation which includes no four phases of intake, compression, expansion, and exhaust.
  • the cycle of operation is achieved as a result of the relative rotation of the inner rotor and outer body, and for this purpose both the inner rotor and outer body may rotate at different speeds, but preferably the inner rotor rotates While the outer body is stationary.
  • One embodiment of the rotating combustion engine which has been successfully used in practice is an engine in which the multi-lobed inner surface of the outer body has substantially the geometric form of an epitrochoid.
  • the shape of the inner rotor illustrated in the drawings resembles in its general configuration the shape of a theoretical rotor having the maximum size which the epitrochoidal inner surface of the outer body can accommodate.
  • the rotor need not have this maximum size but could be cut back between its apex portions, e.g., to reduce the compression ratio of the engine.
  • the shape of the rotor is polygonal or has the general configuration of a polygon having convex or concave curved arcuate sides, the apex portions of the rotor being in continuous contact with the inner wall or surface of the stator.
  • the inner surface of the outer body defines a two-lobed epitrochoid
  • the rotor or inner body has three apex portions and is generally triangular in cross section but has curved or arcuate sides. It is not intended that the invention be limited, however, to the form in which the inner surface of the outer body approximates a two-lobed epitrochoid and the inner body or rotor has only three apex portions.
  • the inner surface of the outer body may be substantially an epitrochoid having one less lobe than the rotor or inner body has apex portions.
  • an end face of the rotor is provided with an internally-toothed ring gear which is concentric with the bearing bore of the rotor.
  • An externally-toothed or pinion gear of smaller diameter than the rotor gear is secured to the outer body and adapted to fit into and mesh with the teeth of the rotor gear as the rotor makes its revolutions relative to the outer body during operation of the engine.
  • the bearing bore of the rotor accommodates a shaft eccentric upon which the rotor is rotatably mounted, and the shaft is coaxial with the outer body cavity.
  • the rotor is constructed of an aluminum alloy (or other lightweight metal alloy having good heat conducting properties) and two metal sleeves of steel or similar material are fitted into the bearing bore :of the rotor and lie between it and the eccentric when the engine is assembled.
  • the inner of these two sleeves also carries the rotor ring gear, and this gear is registered or indexed against relative rotation with respect to the rotor by a series of radial lugs or splines which fonrn a part of the iner sleeve and ring gear, and the radial lugs of ⁇ the rotor gear are in cooperative engagement with complementary radial lugs or splines on the aluminum rotor itself adjacent to its bearing bore.
  • This invention also provides a special notor bearing construction which accommodates the higher operating temperature and higher thermal coetficient of expansion of the aluminum alloy rotor compared to the steel shaft eccentric without excessive bearing clearance; this construction may be referred to as a semi-floating bearing.
  • the outer steel sleeve has a sufficiently tight shrink-fit into the rotor bore that it will remain tight at operating temperatures in spite of the greater thermal expansion of the aluminum alloy rotor compared to that of the outer steel sleeve.
  • the inner steel sleeve also has a shrink-fit into the outer steel sleeve, but it is a considerably less severe shrinlofit than that of the outer sleeve in the aluminum alloy rotor. Even at operating temperature the outer sleeve and aluminum alloy rotor continue to act as one assembly, but at an intermediate temperature before the operating temperature is reached the inner sleeve separates from the outer sleeve, attains its free dimensions,
  • This semifioating hearing design provides much less variation in effective bearing clearance between bearing and eccentric from room temperature to operating temperature than can be provided with a single steel sleeve permanently shrunkfit into an aluminum rotor.
  • the present invention by insuring that the effective bearing clearance between rotor and eccentric remains small in spite of differential thermal expansion be ween these parts, provides the following positive, advanta eous, and beneficial results:
  • the instant invention in making it possible to use a lightweight metal alloy rotor mounted for rotation upon the eccentric of a steel shaft, permits the accomplishment of the following beneficial and advantageous results:
  • the outer body is constructed of a lightweight metal alloy having high heat conducting properties, such as an aluminum alloy
  • the rotor itself must also be constructed of such a lightweight metal alloy to pro ide a continuously small clearance between the end daces of the rotor and the end walls of the outer body. If the outer body were constructed of such a lightweight metal alloy, and the rotor were not similarly constructed, the end walls of the outer body, upon thermal expansion, would grow away from the rotor and provide too great a clearance between the rotor and the outer body to permit elfective sealing between these two components.
  • the present invention by providing means to index or register the rotor gear against rotation relative to the rotor prevents mechanical interference between the rotor and outer body which might otherwise occur because of improper or faulty indexing of the rotor gear with respect to the rotor.
  • Another object of the present invention is to provide means for minimizing the variation of bearing clearance between shaft eccentric and rotor resulting from difierential thermal expansion as the rotary mechanism is elevated from ambient to operating temperature.
  • Another object of the present invention is to provide for a rotor bearing construction that at operating temperature will yield a bearing clearance with the shaft eccentri that is a function only of the bearing free diameter wi" its tolerance band and the eccentric diameter with its tolerance band.
  • Another object of the present invention is to provide means which at operating tem erature will virtually eliminate residual shrink-fit shape distortion or deformation of the rotor bearing caused by the shape of the rotor itself when the mechanism is cold.
  • Another object of the instant invention is to provide a shrink-fit double sleeve for a lightweight metal alloy rotor which has a shrink-fit consistent with the allowable stress properties of aluminum, which can be ssembled on the shaft eccentric when cold, and which will provide the desired bearing clearance between rotor and eccentric at operating temperature and keep this clearance small enough to preserve an oil film on the hearing at all times.
  • Another object of this invention is to provide means to properly circumferentially locate the rotor gear and to index it against rotation with respect to the rotor.
  • a further object of the instant invention is to provide means for registering the rotor gear against relative rotation with respect to the rotor which will maintain substantially the same desired circumferential clearance between the registering means and the rotor when the mechanism is at its operating temperature as exists when the mechanism is cold.
  • a still further object of the instant invention is to provide a sealing means to prevent oil leakage at the rotor end face from the operating clearance between the ou e and inner sleeves which sealing means will also ac ommodate for axial growth of the aluminum rotor relative to the steel inner sleeve as the temperature of the mechanism is elevated.
  • a fundamental purpose of the present invention is to provide a means of restricting the growth of the caring clearance from ambient temperature to operating temperature to as small an increase as possible and at the same time provide a means which will register the rotor against rotation with respect to the rotor gear.
  • the shaft eccentric and both the inner and outer sleeves are of steel. If the eccentric and sleeves are fairly close to the same temperature (which they normally are) and if the sleeves were not restricted, restrained, or forced to grow in any other manner than as steel itself would grow, the double sleeve and the eccentric would maintain a fairly constant clearance as the mechanism warms up to operating temperature.
  • the inner diameter of the shrinkit double steel sleeve will grow faster than a free steel sleeve would.
  • the inner sleeve becomes a free member, attains its free dimensions, and separates itself from the outer steel sleeve and aluminum rotor at some intermediate temperature on route between ambient temperature and operating temperature, it will grow from that temperature on as a free steel member.
  • the circumferential clearance between the aluminum alloy and steel radial splines is designed to be of suflicient magnitude that it will not restrict the radial floating action of the floating hearing or inner sleeve at operating temperature.
  • the shrink-fit between the outer sleeve and the aluminum rotor is adequately severe to be preserved at operating temperature so that the outer sleeve and aluminum rotor always act as a single component or assembly and have no clearance between them.
  • the present invention includes means for minimizing variations in bearing clearance at all temperatures between ambient and operating temperatures which is caused by different rates of thermal expansion for the rotor and eccentric and also includes means for maintaining the relative rotational position of the rotor with respect to its floating bearing and gear at elevate temperatures.
  • the invention consists in the novel parts, constructions, arrangements, combinations, and improvements shown and described.
  • FIG. 1 is a side elevation of the mechanism with the end wall of the outer body removed to show the rotor positioned within the outer body. Portions of the rotor and outer body are shown partially in section, and running bearing clearances and clearances between splines are shown and exaggerated for clarity;
  • FIG. 2 is a central vertical section of the mechanism taken along the line 2-2 of FIG. 1 and in which the eccentric is shown only partially in section; running clearances are shown in an exaggerated manner for clarity;
  • H6. 3 is a partial central vertical section of the mechanism in which, for clarity, a number of the parts are horizontally displaced;
  • FIG. 4 is a diagrammatic view of the built-up rotor as an aid in describing the function of the radial splines in registering the rotor bearing and gear against relative rotation with respect to the rotor; running clearances are shown in an exaggerated manner for clarity;
  • FIG. 5 is a schematic diagram to help clarify the function of the semi-floating bearing in maintaining an almost constant clearance between rotor and eccentric as the mechanism warms up from ambient to operating temperature.
  • FIG. 1 a generally triangular rotor having arcuate sides is eccentrically supported for rotation within an outer body 12.
  • the outer body 12 is fixed or stationary, a practical and useful form of the invention may be constructed in which both the outer body and rotor are rotar but the eccentric is stationary; in this latter form of the invention the powershaft is driven directly by rotation of the outer body and the inner rotor rotates relative to the outer body.
  • a still third form of the invention is possible in which the inner body or rotor is stationary, and the outer body and eccentric are rotatable.
  • the rotor 1t rotates on an axis 14 which is eccentric from and parallel to the axis 16 of the curved inner surface of the outer body 12.
  • the curved inner surface 18 of the outer body 12 is substantially an epitrochoid in geometric shape and includes two arched lobe-defining portions, or lobes.
  • An intake port 219 is arranged to communicate with one lobe of the epitrochoidal inner surface 18, and an exhaust port 22 is arranged to communicate with the other lobe.
  • a line which connects these two points of least radius and passes through the center of the epitrochoid is designated its minor axis 24.
  • the epitrochoid has two points of greatest radius, and a line connecting these two points and passing through the center of the epitrochoid is designated its major axis 26.
  • the minor axis 24 divides the epitrochoid into two halves.
  • the half or lobe which communicates with the exhaust port may be called the exhaust lobe and the half or lobe which communicates with the intake port may be called the intake lobe.
  • the generally triangular shape of the rotor it corresponds in its general configuration to the maximum profile or" the rotor which permits interference iree rotation of the rotor llil with respect to the outer body 12.
  • the outer body 12 which is stationary in this embodiment, comprises two end walls 23 an 3t and a peripheral wall 32 interconnecting these end walls.
  • a crankshaft 34 is rotatably supported by the end walls 28 and 3b of the outer body 12 by means of conventional bearings, and the axis of the crankshaft 34 is coincident with the axis 16 of the outer body 12.
  • An eccentric 36 is rigidly mounted on and forms an integral part of the crankshaft 34; the axis of the eccen tric 36 is eccentric from and parallel to the crankshaft axis 16. T he rotor llfl is rotatably supported upon the eccentric 36, and the central axis of the eccentric 36 is coincident with the axis 14- of the rotor lib.
  • the rotor It includes three apex portions 38 which carry radially movable sealing members
  • the sealing members 4t? are in substantially continuous sliding, gas-sealing contact with the inner surface 18 of the outer body 12 as the rotor it rotates within and relative to the outer body 12.
  • variable volume working chambers 42 are formed between the outer peripheral working faces of the rotor 10 and the inner surface 18 of the outer body 12.
  • the rotation of the rotor relative to the outer body is counterclockwise and is indicated by an arrow.
  • a spark plug is mounted in the peripheral wall 32 of the outer body 12, and at the appropriate time in the engine cycle, the spark plug as provides ignition for a compressed combustible mixture which upon expansion drives the rotor in the direction of the arrow.
  • the eccentri ity (e) of the rotor axis 14 from the outer body axis 16 acts as a crank arm or moment arm to convert the energy of the expanding gases into torque on the crankshaft 34.
  • this means comprises a pair of steel sleeves and 52.
  • the bearing bore 43 of the aluminum rotor 10 is lined with a steel outer sleeve 5d which has a severe shrinka lit in the aluminum bearing bore The shrinlofit between the bearing bore 43 and the outer sleeve is so severe that it approaches the limit of elasticity of the aluminum rotor 1% beyond which plastic deformation of the aluminum would occur.
  • the outer sleeve is cooled below the ambient temperature, and the aluminum rotor 1% is raised above the ambient temperature to permit assembly of the interference-fit rotor 1d and sleeve
  • the cold outer sleeve 59 is placed in the bearing bore 43 of the hot aluminum rotor, so that when the parts return to the ambient temperature a severe shrink-fit results.
  • This shrink-fit is sufiicient-ly severe so that even when operating temperature of the mechanism is reached, the outer sleeve 5% keeps its shrink-fit in the bearing bore 43 and the outer sleeve 5% and aluminum rotor 1% ⁇ act as one assembly throughout the full range of temperatures of the mechanism.
  • the inner steel sleeve 52 In assembly of the rotor lit), the inner steel sleeve 52 also has a shrink-fit in the rotor it or more precisely, the inner sleeve 52 has a shrink-fit in the inner diameter of the outer sleeve 5'8.
  • the shrink-fit between the inner sleeve 52 and the outer sleeve 5 however, is considerably less severe than the shrink-fit between the outer sleeve 5t and the bearing bore 43 of the aluminum rotor ill but is suificient to ensure that at ambient temperature the inner sleeve 52, the outer sleeve 5%, and the aluminum rotor 10, will all act as one assembly.
  • the relatively slight shrink-fit of the inner sleeve 52 and the outer sleeve 54% is overcome by the radially outward expansion of the assembly of the outer sleeve 59 and the aluminum rotor 19.
  • the inner sleeve 52 attains its free dimensions for the intermediate temperature at which the separation occurs and continues to expand thermally as a free steel member, whereas the assembly of outer sleeve and rotor continues to expand thermally at a greater rate, which is some rate of expansion intermediate between the rates for free steel and free aluminum.
  • the inner sleeve 52 acts as a true floating bearing, since at those temperatures there will be a bearing clearance between the inner sleeve 52 and the outer sleeve 5% as well as the normal hearing clearance between the inner sleeve 52 and eccentric 36 which exists even at room temperature.
  • means are provided to register the rotor gear against relative rotation with respect to the rotor.
  • An internally-toothed ring gear 54 is secured to the inner sleeve 52, and as embodied, the means for registering or indexing the rotor gear 5 3- against rotation relative to the rotor it? comprise radially faced splines or lugs 56 which are secured to the inner sleeve 52 and mesh with correspondingly shaped splines se.
  • the inner sleeve or ring gear 54 is in mesh or engagement with an externally-toothed or pinion gear 53 which is rigidly attached to the outer body 12.
  • the gear ratio between the rotor gear 54 and outer body gear 58 is 3 :2, so that for every revolution of rotor in about its own axis 14, the crankshaft 34 rotates three times in the same direction.
  • the purpose of the gearing between the rotor and the outer body is to register or index the rotor in its position within the outer body relative to the outer body, and to relieve the load on the apex portions 38 and sealing members 41? of the rotor, since these parts would otherwise bear the load of determining the registration of the rotor and the outer body.
  • the gearing thus, does not function to impart torque to the crankshaft 34, which is accomplished through the eccentric 36.
  • the radial splines as of the inner sleeve 52 mate with complementary radially faced splines or lugs 6b in the aluminum rotor lil itself.
  • both the inner sleeve splines 56 and the rotor splines 66 are radially faced and since both the inner sleeve 52 and rotor it? grow thermally along radial lines, i.e., their thermal growth is proportional to the radius at the point or" growth, the circumferential clearance between the inner sleeve splines 56 and the rotor splines 6% ⁇ remains constant regardless of temperature.
  • an adjustable nut 62 having a radially outward extending flange 64 is threadedly engaged with the outer diameter or" the inner sleeve 52 (see FIG. 2) on the axial end of the inner sleeve 52 0pposite the end which carries the ring gear 5
  • the purpose of the adjustable nut 62 with its fiange 64 and the shoulder as of the inner sleeve 52 is to seal lubricating oil from leaking out along the end face 66 of the rotor lb from the clearance between the inner sleeve 52 and the outer sleeve 56 when the engine is operating.
  • This steel flange 64- is suiiiciently thin to permit resilient deflection to allow for axial thermal expansion or" the outer sleeve 5% and aluminum rotor 16.
  • Channels 45 are cut in the working faces 44 of the rotor ll) both to r Jerusalem the compression ratio or" the engine and to permit the flow of combustible gases between the intake lobe and exhaust lobe of the working chamber 42 when it is located in firing position opposite the spark plug 46, as shown in FIG. 1.
  • FIG. 4 is a diagrammatic or schematic view of the built-up aluminum rotor, eccentric, and crankshaft 34. This view shows the running clearances, exaggerated for clarity, between the parts depicted and its purpose is to clarify the need for maintaining a relatively constant circumferential clearance between the inner sleeve splines 56 and rotor splines 6%.
  • inner sleeve splines 56 For simplification, only four of the inner sleeve splines 56 are shown in solid line. These radially-faced inner sleeve splines 56 dovetail or mesh with radially-faced splines 61 in the rotor 10. As shown in FIG. 4, when the mechanism is at operating temperature, the inner sleeve 52 acts as a floating hearing so that there is a finite clearance between the inner sleeve 52 and the outer sleeve 5% as well as the usual bearing clearance between the inner sleeve 52 and the eccentric 36.
  • the diametral running clearance between the inner and outer sleeves and the diametral running clearance between the inner sleeve and the eccentric are approximately equal, and each of these clearances is equal to about 0.2% of the diameter of the eccentric.
  • the floating action of the bearing presents a problem in achieving registraiton of the inner sleeve against rotation with respect to the rotor by means of the interlocking splines 56 and 61;: (FIG. 1), since when a load or force is applied to the rotor, as when a Working face is undergoing the expansion phase of the engine cycle, the floating bearing clearance will essentially disappear, or be taken up, on the side of the rotor adjacent to the direction from which the force is applied.
  • this force is represented by the arrow designated P which is shown acting radially against the working face 44. It is apparent from FIG. 4 that there will be a small amount of motion when the bearing clearances are taken up as the load or force is applied to the working face 44. This load or force F is reacted on the diameter of the eccentric 36, but the rotor must move sub stantially through the bearing clearances before the load can be efiectively reacted. In other words, there is some relative radial motion between the rotor 18, the inner sleeve 52, and the eccentric 36.
  • a finite clearance equal to or greater than the total radial bearing clearance is provided between the splines 56 and the corresponding rotor splines as; thus, it the total radial bearing clearance is 0.21% of the diameter of the eccentric, the clearance between the radially aligned gear faces of the splines 56 and the rotor splines 6% must be equal to or greater than 0.2% of the eccentric diameter.
  • the provision of this constant circumferential clearance between gear splines 56 and rotor splines 6% permits the entire radial bearing clearance to be taken up or eliminated without placing any stress on the splines, and permits the load to be reacted directly on the eccentric 36.
  • FIG. is a sc ematic or diagrammatic representation which shows how the objects of the present invention are achieved by a built-up aluminum alloy rotor construction (comprising an aluminum alloy rotor itself, a steel outer sleeve, and a steel inner sleeve) in cooperation with an eccentric upon which the rotor is mounted for rotation.
  • the horizontal or X axis of PH ⁇ . 5 is a plot of the temperature of the rotor in degrees Fahrenheit, and the vertical or Y axis is a plot of the relative thermal expansion or increase in size of the principal parts of the built-up rotor and the eccentric due to temperature elevation.
  • the lines representing the thermal expansion of the free aluminum alloy rotor and free outer sleeve are hypothetical, since, throughout the operating temperature range, the rotor and outer sleeve act as one assembly, and the line representing their combined thermal expansion is a line having a slope intermediate between the slope of the thermal expansion line for free aluminum alloy (rotor) and the slope of the thermal expansion line for free steel (sleeve).
  • the assembled aluminum alloy rotor has a greater inner diameter than the free rotor would have, and conversely the assembled steel outer sleeve has a smaller outer dimeter than the free outer sleeve would have.
  • the inner steel sleeve also is not free but has a shrink-fit in the outer steel sleeve.
  • the slope of the equation representing thermal expansion of the complete assembly from ambient temperature (76 F.) to approximately F. is represented by the dashed line designated rotor-t-outer sleeve-i-inner sleeve assembly.
  • the shrink-fit of the inner sleeve within the outer sleeve is overcome and the inner sleeve from that temperature on acts as a free steel member and follows the normal thermal expansion curve for steel.
  • the loss of the inner sleeve from the assembly at 160 F. decreases the influence of the steel on the thermal expansion curve for the assembly and, aceo-rdingh changes the slope of the curve r presenting the assembly, as can be seen in PEG. 5.
  • the curve of thermal expansion for the eccentric follows the normal curve for a free steel member throughout the temperature range.
  • the clearance at ambient temperature between the eccentric and the inner diameter of the inner sleeve is designated on FIG. 5, and the clearance at 450 F. is also designated.
  • a comparison of these two clearances shows that their diiference is only slight, and thus one of the primary objects of the invention is achieved in minimiz ing the variation of the clearance between the inner sleeve and eccentric throughout the operating temperature range.
  • Ki. 5 also shows how the clearance between the inner and outer sleeves grows from a shrink-fit and no clearance at room temperature, to the beginning of a finite clearance at 160 F, and finally to a clearance between the sleeves at operating temperature which is approximately equal to the operating clearance between the inner sleeve and the eccentric.
  • the sum of these two operating clearances is equal to the floating bearing clearance or total radial bearing clearance, and two clearances, one on each side of the inner sleeve, give the inner sleeve the possibility of eccentric radial movement and also determine the limits of that movement.
  • the rotor and outer sleeve assembly continuously act as one piece or one assembly throughout the range of temperatures between ambient temperature and operating temperature.
  • the inner sleeve attains its free detailed dimensions and becomes a free floating steel member, so that the only tolerance influencing the clearance between it and the eccentric is the tolerance on its inner diameter and the tolerance band of the eccentric. All of the effects of the tolerance on the outer diameter of the inner sleeve, the tolerance band of the outer sleeve, and the tolerance band of the aluminum rotor that would normally contribute to the effective tolerance of the inner diameter of the inner sleeve when the mechanism is cold, are eliminated, because the inner sleeve is free.
  • the operating clearance thus, is much less aflected.
  • the operating or running clearance is, therefore, a function of only two tolerances, the tolerances on the outer diameter of the eccentric and the inner diameter of the inner sleeve.
  • a rotor for rotary mechanisms having an axis and supported for rotation upon and relative to an eccentric portion of a member, the rotor having an inner bore; an outer sleeve having a shrink-fit into the inner bore with sutflcient tightness to maintain the shrink-fit at all temperatures between the ambient temperature and the operating temperature of the rotor; an inner sleeve bearing shrunk-fit into the inner diameter of the outer sleeve with a shrink-tit that is less tight than the shrink-fit of the outer sleeve into the inner bore; whereby the shrink-fit of the inner sleeve bearing is overcome at an intermediate temperature between the ambient temperature and the operating temperature of the rotor so that the inner sleeve bearing acts as a floating bearing between the eccentric portion and outer sleeve above the intermediate temperature.
  • a rotor for rotary mechanisms having an axis and suppor ed for rotation upon and relative to an eccentric portion or" a member; bearing means between the rotor and eccentric portion for dividing increases in bearing clearance between the rotor and eccentric portion due to differential thermal dimensional changes of the rotor relative to the eccentric portion; the bearing means including a sleeve having a shrink-fit in the inner perimeter of the rotor sufficiently tight to remain in tight contact with the rotor at maximum operating temperatures; and an inner b adjacent to the eccentric portion and having a sk-fit within the sleeve that is less tight than the fit of me sleeve in the rotor; the shrink-fit of the inner bearing in the sleeve being overcome at an intermediate temperature between the ambient temperature and the operating temperature of the rotor; whereby the inner bearing acts as a floating bearing between the eccentric portion and sleeve above the intermediate temperature and divides the total clearance between the hearing and sleeve into two clearances: (l) a clearance between
  • the invention as defined in claim 6, that also includes an internally-toothed gear rotatable with the rotor and secured to the inner bearing; the means for securing the internally-toothed gear to the inner bearing comprising external splines on the internally-toothed gear and internal splines on the rotor in interlocking engagement with the external splines with clearance both radially and circumterentially between the external and internal splines; the external and internal splines registering the rotor and internally-toothed gear against relative rotation regardless of differential thermal expansion between the rotor and the internally-toothed gear as the rotor varies in temperature between the ambient temperature and its operating temperatures.

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  • Engineering & Computer Science (AREA)
  • General Engineering & Computer Science (AREA)
  • Mechanical Engineering (AREA)
  • Chemical & Material Sciences (AREA)
  • Combustion & Propulsion (AREA)
  • Rotary Pumps (AREA)
  • Pistons, Piston Rings, And Cylinders (AREA)
US26233A 1960-05-02 1960-05-02 Rotor and bearing construction for rotary mechanisms Expired - Lifetime US3111261A (en)

Priority Applications (5)

Application Number Priority Date Filing Date Title
US26233A US3111261A (en) 1960-05-02 1960-05-02 Rotor and bearing construction for rotary mechanisms
DEC23918A DE1158750B (de) 1960-05-02 1961-04-19 Rotationskolbenbrennkraftmaschine
CH472161A CH383070A (de) 1960-05-02 1961-04-21 Rotationskolbenmaschine, insbesondere -Brennkraftmaschine
BE603137A BE603137A (fr) 1960-05-02 1961-04-27 Machine et, en particulier, moteur à combustion interne, à piston rotatif
GB15683/61A GB939704A (en) 1960-05-02 1961-05-01 Rotor and bearing construction for rotary internal combustion engines, fluid motors, pumps, compressors and the like

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US26233A US3111261A (en) 1960-05-02 1960-05-02 Rotor and bearing construction for rotary mechanisms

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BE (1) BE603137A (fr)
CH (1) CH383070A (fr)
DE (1) DE1158750B (fr)
GB (1) GB939704A (fr)

Cited By (14)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US3240423A (en) * 1965-05-14 1966-03-15 Curtiss Wright Corp Composite shaft for rotary combustion engine
US3280803A (en) * 1963-09-27 1966-10-25 Sabet Huschang Rotary internal combustion engine
US3297240A (en) * 1965-04-19 1967-01-10 Toyo Kogyo Company Ltd Rotary piston mounting mechanism
US3333763A (en) * 1966-02-02 1967-08-01 Nsu Motorenwerke Ag Sealing arrangement for rotary engines
US3356291A (en) * 1965-10-06 1967-12-05 Thomas W Kennedy Rotary piston machine
US3369740A (en) * 1966-05-04 1968-02-20 Kloeckner Humboldt Deutz Ag Rotary piston internal combustion engine, especially circular piston internal combustion engine
US3383936A (en) * 1967-02-13 1968-05-21 Curtiss Wright Corp Light-weight rotor and gear assembly for rotary mechanisms
US3440929A (en) * 1966-11-10 1969-04-29 Sachsenring Automobilwerke Rotary piston type of combustion engine
US3804560A (en) * 1971-10-08 1974-04-16 Daimler Benz Ag Oil shield arrangement for rotary piston engine
US3963387A (en) * 1975-05-01 1976-06-15 Curtiss-Wright Corporation Rotary engine with self-centering rotor gear
US3989423A (en) * 1974-07-04 1976-11-02 Audi Nsu Auto Union Aktiengesellschaft Piston of light metal for a rotary piston combustion engine
US4032268A (en) * 1974-03-07 1977-06-28 Wankel Gmbh Rotary piston engine
US20120051915A1 (en) * 2010-08-31 2012-03-01 Mitsubishi Heavy Industries, Ltd. Planetary gear train with improved bearing structure and manufacture method of the same
US20130114999A1 (en) * 2010-06-30 2013-05-09 Thomas Östling Driver for torque and rotation transfer from a rotational chuck to a drill steel

Families Citing this family (2)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
DE1551098A1 (de) * 1966-11-11 1970-02-05 Daimler Benz Ag Rotationskolben-Brennkraftmaschine
CN113175480B (zh) * 2021-05-18 2023-01-06 杰锋汽车动力系统股份有限公司 一种氢气循环泵轴承结构

Citations (8)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US1456171A (en) * 1922-04-17 1923-05-22 Millard F Woodward Internal-combustion engine
US1498806A (en) * 1922-12-11 1924-06-24 Jr Joseph H Oberman Pump cylinder
US1529911A (en) * 1922-12-18 1925-03-17 Independent Pneumatic Tool Co Removable wear bushing for cylinders
GB557902A (en) * 1942-09-23 1943-12-09 Harold Kershaw Improvements in rotary piston machines
GB583035A (en) * 1943-08-20 1946-12-05 Bernard Maillard A rotary machine generating variable volumes
FR1125876A (fr) * 1954-06-09 1956-11-09 Nsu Werke Ag Machine à piston rotatif à axe intérieur
US2947290A (en) * 1957-11-18 1960-08-02 Nsu Werke Ag Heat generating rotary internal combustion engine
US2988065A (en) * 1958-03-11 1961-06-13 Nsu Motorenwerke Ag Rotary internal combustion engine

Patent Citations (8)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US1456171A (en) * 1922-04-17 1923-05-22 Millard F Woodward Internal-combustion engine
US1498806A (en) * 1922-12-11 1924-06-24 Jr Joseph H Oberman Pump cylinder
US1529911A (en) * 1922-12-18 1925-03-17 Independent Pneumatic Tool Co Removable wear bushing for cylinders
GB557902A (en) * 1942-09-23 1943-12-09 Harold Kershaw Improvements in rotary piston machines
GB583035A (en) * 1943-08-20 1946-12-05 Bernard Maillard A rotary machine generating variable volumes
FR1125876A (fr) * 1954-06-09 1956-11-09 Nsu Werke Ag Machine à piston rotatif à axe intérieur
US2947290A (en) * 1957-11-18 1960-08-02 Nsu Werke Ag Heat generating rotary internal combustion engine
US2988065A (en) * 1958-03-11 1961-06-13 Nsu Motorenwerke Ag Rotary internal combustion engine

Cited By (15)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US3280803A (en) * 1963-09-27 1966-10-25 Sabet Huschang Rotary internal combustion engine
US3297240A (en) * 1965-04-19 1967-01-10 Toyo Kogyo Company Ltd Rotary piston mounting mechanism
US3240423A (en) * 1965-05-14 1966-03-15 Curtiss Wright Corp Composite shaft for rotary combustion engine
US3356291A (en) * 1965-10-06 1967-12-05 Thomas W Kennedy Rotary piston machine
US3333763A (en) * 1966-02-02 1967-08-01 Nsu Motorenwerke Ag Sealing arrangement for rotary engines
US3369740A (en) * 1966-05-04 1968-02-20 Kloeckner Humboldt Deutz Ag Rotary piston internal combustion engine, especially circular piston internal combustion engine
US3440929A (en) * 1966-11-10 1969-04-29 Sachsenring Automobilwerke Rotary piston type of combustion engine
US3383936A (en) * 1967-02-13 1968-05-21 Curtiss Wright Corp Light-weight rotor and gear assembly for rotary mechanisms
US3804560A (en) * 1971-10-08 1974-04-16 Daimler Benz Ag Oil shield arrangement for rotary piston engine
US4032268A (en) * 1974-03-07 1977-06-28 Wankel Gmbh Rotary piston engine
US3989423A (en) * 1974-07-04 1976-11-02 Audi Nsu Auto Union Aktiengesellschaft Piston of light metal for a rotary piston combustion engine
US3963387A (en) * 1975-05-01 1976-06-15 Curtiss-Wright Corporation Rotary engine with self-centering rotor gear
US20130114999A1 (en) * 2010-06-30 2013-05-09 Thomas Östling Driver for torque and rotation transfer from a rotational chuck to a drill steel
US10174565B2 (en) * 2010-06-30 2019-01-08 Epiroc Rock Drills Aktiebolag Driver for torque and rotation transfer from a rotational chuck to a drill steel
US20120051915A1 (en) * 2010-08-31 2012-03-01 Mitsubishi Heavy Industries, Ltd. Planetary gear train with improved bearing structure and manufacture method of the same

Also Published As

Publication number Publication date
GB939704A (en) 1963-10-16
CH383070A (de) 1964-10-15
DE1158750B (de) 1963-12-05
BE603137A (fr) 1961-08-16

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