CN111247316A - Synchronous belt driving system - Google Patents

Abstract

A synchronous belt drive system comprises: a synchronous belt having a tensile cord made of a high modulus fiber such as glass, carbon, PBO or aramid; a drive sprocket and at least one driven sprocket, at least one of the sprockets being an oblong sprocket; and a tensioner having: a base having an axially extending cylindrical portion with a radially outer surface and a receiving portion; an eccentric arm pivotally engaged with the radially outer surface; a torsion spring disposed within the receiving portion, the torsion spring applying a biasing force to the eccentric arm; and a pulley journalled to the eccentric arm. Preferably, none of the eccentric arms, pulleys, or torsion springs are axially displaced from each other along the axis a-a. The obround sprocket has a toothed surface and at least one linear portion disposed between two arcuate portions, the arcuate portions having a constant radius, the linear portion having a predetermined length.

Description

Synchronous belt driving system

Technical Field

The present invention relates generally to a synchronous belt drive system, such as used in an internal combustion engine, comprising a timing belt, an automatic mechanical tensioner, a drive pulley and one or more driven pulleys, at least one of which is oblong, and more particularly to a belt drive system adapted to run in oil and having a narrower package width than known systems.

Background

Most synchronous belt drive systems for engines are mounted on the front of the engine, outside the engine block. In such so-called "dry" belt applications, the timing belt may operate without lubrication. As a result of this approach, the drive shaft and camshaft(s) must exit the engine block through oil seals that may be susceptible to leakage over time. The benefit is that the width of the tape drive can be accommodated relatively easily in the engine compartment, i.e. relatively without limitation to the drive system width. However, some engine compartments have limited space and it is desirable to reduce the weight of the vehicle, so a narrower dry belt drive system that handles the same drive load, tension and timing requirements with reduced system weight and width is desirable. In this context, "width" refers to the dimension of the drive system or drive member in the axial direction (i.e., the direction perpendicular to the plane of the drive when viewed in the layout of the drive).

Timing chains for internal combustion engines are typically installed inside the engine where they are easily lubricated. Another engine timing approach is to install a synchronous belt drive system inside the engine block, similar to the design of the timing chain drive. This approach may reduce the number of seals required on the shaft. This approach may require more restrictions on system width to avoid the expense of oversized engine blocks when matching the width of the timing chain drive for the same application. In addition, such so-called "wet" belt applications require belt materials having excellent resistance to oil and other engine fluids. Accordingly, what is desired is a narrower wet belt drive system that can directly replace the timing chain system, operate in contact with oil while handling the same drive loads, tensions, and positive timing requirements with reduced system weight.

As an example, one known commercial belt-in-oil (belt-in-oil) timing drive system has a belt that is about 18mm wide, a circular pulley that is about 19-23mm wide, and a conventional tensioner that occupies a width of about 35-38 mm. The peak timing error is considered to be typically about 2 peak-to-peak. The total mass of the system including the VVT components is reported to be about 2500g, and the total mass of the system without the VVT components is reported to be about 1700 g.

U.S. Pat. No. 9,927,001 (Dayco Europe SRL) is representative of the art. Therein, a belt life test is described using a 19-mm wide belt running in contact with oil.

It is desirable to reduce the total system width to less than 18mm, i.e., less than half the typical current system width. For this purpose, at least slightly narrower belts and pulleys and very narrow tensioners are required. This would be an inconvenient task as reducing the belt width would be expected to increase belt tensile strain, contact pressure, and tooth deflection, which in turn may increase timing error and accelerate belt degradation, thereby reducing timing system performance and life expectancy. It is desirable to reduce the system width to 20mm or less and reduce weight equally while reducing timing error by about 50% and without losing belt life.

Disclosure of Invention

The present invention relates to systems and methods that provide a narrower wet or dry belt timing drive system that can handle typical drive loads, tensions, and timing requirements with reduced system weight and width.

The invention relates to a system having a high modulus timing belt, an oblong sprocket, and a narrow automatic mechanical tensioner. The high modulus timing belt preferably has tensile cords made of high modulus fibers such as high modulus glass fibers, carbon fibers, PBO or aramid, and mixtures thereof. The phase and amplitude of the obround sprocket are selected to reduce timing error to less than 1.5 peak-to-peak, preferably less than 1.0 peak-to-peak. The obround sprocket has a toothed surface and at least one linear portion disposed between two arcuate portions of constant radius, the linear portion having a predetermined length. The tensioner damping and tension are selected to maintain slack side belt tension preferably in the range of 100N to 600N, preferably with a torsion spring and asymmetric damping. The tensioner may preferably have: a base having an axially extending cylindrical portion with a radially outer surface and a receiving portion; an eccentric arm pivotally engaged with the radially outer surface; a torsion spring disposed within the receiving portion, the torsion spring applying a biasing force to the eccentric arm; and a pulley journalled to the eccentric arm. Preferably, none of the eccentric arms, pulleys or torsion springs are axially displaced from each other along the axis a-a.

The foregoing has outlined rather broadly the features and technical advantages of the present invention in order that the detailed description of the invention that follows may be better understood. Additional features and advantages of the invention will be described hereinafter which form the subject of the claims of the invention. It should be appreciated by those skilled in the art that the conception and specific embodiment disclosed may be readily utilized as a basis for modifying or designing other structures for carrying out the same purposes of the present invention. It should also be realized by those skilled in the art that such equivalent constructions do not depart from the scope of the invention as set forth in the appended claims. The novel features which are believed to be characteristic of the invention, both as to its organization and method of operation, together with further objects and advantages will be better understood from the following description when considered in connection with the accompanying figures. It is to be expressly understood, however, that each of the figures is provided for the purpose of illustration and description only and is not intended as a definition of the limits of the present invention.

Drawings

The accompanying drawings, which are incorporated in and constitute a part of the specification in which like numerals designate like parts, illustrate embodiments of the present invention and together with the detailed description, serve to explain the principles of the invention.

FIG. 1 is a side view of an oblong sprocket.

Fig. 2 is a side view of an alternative obround sprocket.

FIG. 3 is a perspective view of a dual cam, in-line four cylinder, four stroke gasoline engine.

FIG. 4 is a perspective view of a single cam, in-line four cylinder, four stroke diesel drive engine having a fuel pump driven at the rear of the camshaft.

FIG. 5 is a perspective view of a single cam, four cylinder, four stroke diesel driven engine having a fuel pump incorporated in the synchronous belt drive system.

FIG. 6 is a schematic diagram of a dual cam, four cylinder, four stroke gasoline powered engine.

Fig. 7 represents typical total load characteristics of a driven sprocket for a four cylinder, four stroke diesel engine, which contains extracted 1.5 and 2 order curves.

Fig. 8 represents the 2 nd order load characteristic of the drive sprocket for a four cylinder, four stroke engine.

Fig. 9 represents the 1.5 order load characteristic of the drive sprocket (or other device causing 1.5 order) of a four cylinder, four stroke common rail diesel engine with a 3-piston fuel pump.

Figure 10 is a series of curves representing the stress/strain relationship of the timing belt.

Fig. 11 is a series of curves showing the effect of phasing/phase shift of an obround sprocket on engine dynamics of the system in fig. 6.

Fig. 12 is a graph showing angular vibration characteristics at the camshaft of the engine shown in fig. 6 before and after application of the obround sprocket.

FIG. 13 is a graph showing the tight side tension profile of the engine shown in FIG. 6 before and after application of the obround sprocket.

FIG. 14 is a graph of vibration angle versus crankshaft speed.

Fig. 15 is a graph of vibration angle versus crankshaft speed for an intake cam.

Fig. 16 is a graph of oscillation angle versus crankshaft speed for an exhaust cam.

FIG. 17 is a graph of angular displacement versus crankshaft speed for an intake cam.

FIG. 18 is a graph of angular displacement versus crankshaft speed for an exhaust cam.

FIG. 19 is a graph illustrating the effect of phasing of an obround sprocket on individual camshaft timing errors.

Fig. 20 is a graph showing the effect of phasing of an oblong sprocket with respect to individual camshaft timing errors in the case of a standard high modulus belt.

Fig. 21 is a graph showing the effect of an oblong sprocket on timing error due to belt width.

FIG. 22 is a graph illustrating the effect of the magnitude of the oblong sprocket on timing error due to eccentricity.

Fig. 23 is an exploded view of the tensioner.

Fig. 24 is a top exploded view of the tensioner.

Fig. 25 is a perspective view of the base of the tensioner.

FIG. 26 is a perspective view of an eccentric arm of the tensioner.

Fig. 27 is a perspective view of a torsion spring of the tensioner.

Fig. 28 is a cross-sectional view of the tensioner.

Fig. 29 is an exploded view of an alternative tensioner.

Fig. 30 is a top view of the alternative tensioner of fig. 29.

Fig. 31 is a cross-sectional view of the alternative tensioner of fig. 29.

Fig. 32 is a side view of an alternative tensioner.

Fig. 33 is a perspective view of the alternative tensioner of fig. 32.

Fig. 34 is a partial view of a timing belt.

Figure 35 is a diagram of a synchronous belt drive system used to test aspects of the present invention.

Detailed Description

The present invention relates to a synchronous belt drive system comprising a high modulus timing belt, one or more oblong sprockets, and a narrow design automatic mechanical tensioner. Preferably, the high modulus timing belt has a tensile modulus that is about twice the modulus of a timing belt of similar width but with a conventional fiberglass tensile member. The high modulus timing belt is preferably constructed of a temperature and oil resistant material. Preferably, the amplitude and phase of the obround sprocket are such that the angular displacement timing error between the sprocket and the obround sprocket is less than 1.5 degrees peak-to-peak. The automatic mechanical tensioner may include a base having an axially extending cylindrical portion including a radially outer surface and a receiving portion radially inward of the radially outer surface; an eccentric arm pivotally engaged with the radially outer surface; a torsion spring disposed within the radially inner receiving portion, the torsion spring applying a biasing force to the eccentric arm; and a pulley having a shaft diameter connected to the eccentric arm. Preferably, the eccentric arm, the pulley, and the torsion spring are concentrically arranged such that none of the eccentric arm, the pulley, or the torsion spring is axially displaced from any of the eccentric arm, the pulley, or the torsion spring along the axis a-a, thereby forming a minimum width.

A high modulus timing belt.

Fig. 34 shows a configuration of the timing belt 200. The belt 200 includes teeth 214 on one side (which alternate with a tooth bottom 215) and a smooth back side 220. The body rubber or elastomer includes a tooth rubber 212 and a back rubber 222. The tooth side is covered by a tooth sheath 216 and the back side 220 is covered by a back sheath 224. The repeating length of the teeth is the pitch "P". Tensile member 218 is embedded in the belt body rubber and imparts a high modulus to the belt.

The tooth sheath 216 includes fabric and one or more treatments to enhance one or more belt properties (e.g., adhesion, oil resistance, wear resistance, etc.) and to enhance system performance properties (such as timing error and durability). The fabric treatment may be any suitable treatment known in the art. Likewise, the back jacket 224 may comprise a fabric and one or more treatments that are the same or different than the tooth jacket. Thus, the term "sheath" is used to describe a fabric containing one or more treatments, typically in a form in which it is ready to be assembled into a belt. "Fabric" generally refers to a woven, nonwoven, or knitted material of the blank prior to application of the treatment.

The facing fabric may be any suitable woven, knitted, or nonwoven fabric having suitable stretch, strength, abrasion resistance, temperature resistance, and environmental resistance, depending on the needs of the application. For use in a flow-through tape manufacturing process (defined later), a machine direction extensibility of more than 80% or more than 100% is preferred. For a pre-forming process, very low stretchability may be suitable, or no significant stretchability may be suitable. The fabric may comprise high strength, oil resistant, wear and temperature resistant fibers such as nylon, aramid, PPS, PEEK, polyester, and combinations thereof. Yarns with sufficient stretch can be obtained by any known suitable method including crimping, wrapping an elastic core, orienting on twill, and combinations of these methods.

The back fabric can be any suitable woven, knitted, or nonwoven fabric having suitable stretch, strength, abrasion resistance, temperature resistance (heat or cold) and environmental resistance for the application. Generally, there is no particular requirement for the amount of any stretch for the back fabric, as the back fabric will only lie flat on the back of the belt. To maintain the flexibility of the belt, a certain degree of stretch may be preferred. It has been found that the back fabric can improve cold resistance, thereby reducing back cracking due to repeated cold starts.

Preferred fabric treatments include epoxy or epoxy-rubber treatments as described in U.S. patent publication No. 2014/0080647a1 to Yamada et al, which is incorporated herein by reference, and optionally RFL treatments. Such treatment is intended to improve the wear resistance and the oil and water resistance of the tooth sheath, and to provide such a toothed belt: the toothed belt has satisfactory durability even when used under high-temperature and high-load conditions or in an oil or water environment.

Any suitable rubber composition(s) may be used for the tooth rubber 212 or the back rubber 222. In addition, other rubber layers may be present, such as an adhesive rubber layer in contact with the tensile cords 218, or other layers, as desired. The same or different compounds may be used in the teeth, in the tensile cord layer, on the back side, and elsewhere in the belt as desired. Exemplary rubber compounds for tooth rubber or belt body rubber are described in U.S. patent No. 6,358,171B1 to Whitfield, which is incorporated herein by reference. As described therein, the belt body rubber composition may include a nitrile group-containing copolymer rubber, such as HNBR, and the rubber may include a third monomer that reduces the glass transition of the rubber. The rubber composition may also include about 0.5 to about 50 parts by weight rubber (PHR) of a fibrous reinforcement. Other exemplary rubber compounds for tooth rubber or belt body rubber are described in U.S. patent No. 9,140,329B2, which is incorporated herein by reference. As described therein, the belt body rubber composition may include HNBR or HXNBR rubber, resorcinol, and a melamine compound.

The rubber composition(s) of the belt body may further comprise additional ingredients known in the art such as fillers, plasticizers, antidegradants, processing aids, curatives, adjuvants, compatibilizers, and the like.

The tensile cords 218 for the belt may be any known in the art, but preferably comprise glass fibers, PBO, aramid, carbon fibers, or mixtures of two or more of the foregoing. The tensile cord preferably includes an adhesion treatment having high resistance to oil for use in an oil-wet environment. For example, the bonding treatment may be based on latex or rubber containing nitrile or other oil resistant materials. Preferred tensile cords include carbon fibers, such as disclosed in U.S. Pat. No. 6,945,891 to Knutson, or glass/carbon hybrid cords, such as described in U.S. Pat. No. 7,682,274 to Akiyama et al. Preferred glass fibers for the tensile cord include high strength glass fibers such as K-glass, U-glass, M-glass, or S-glass.

The toothed belt of the present invention can be manufactured according to known methods of manufacturing belts. The most common method is to apply various materials to a grooved mandrel, first a tooth cover sheath, then a tensile cord and body rubber, and finally a back sheath. The mandrel with the ribbon plate is then inserted into a pressurizable housing (which can be heated and pressurized to squeeze the materials together) to flow the rubber into the gullets, pushing the tooth sheath into the shape of the gullets (a process known as "flow-through"). Alternatively, the tooth sheath may be preformed to approximate the shape of the groove and optionally filled with rubber (a "preforming process") prior to or simultaneously with placing the tooth sheath on the mandrel. Other variations of these methods are also possible. A major additional feature for making fabric-backing belts is that the rubber layer must be carefully measured in order to obtain the desired final belt thickness, since the back of the belt cannot be ground to size as for rubber-backing belts.

Rubber compounds generally have such modulus levels: this modulus level contributes significantly to the tooth stiffness and the load capacity of the tooth and timing belt. Also, the tooth cover sheath helps to strengthen the teeth, which also helps with tooth stiffness and load capacity of the belt. Tensile cords typically dominate the tensile properties of the timing belt, such as modulus (or span stiffness) and strength. It has been found that optimizing the combination of belt span stiffness and tooth stiffness by selecting these materials allows narrower belts to have the same system timing error while reducing system loads and minimizing the required strength in the belt. In particular, increasing the span stiffness relative to that of conventional belts with glass fiber cords has the desired effect. In computer simulations, a series of belts with up to twice span stiffness and the same tooth stiffness have the same system timing error, but reduce the system maximum belt tension, maximum belt effective tension, and maximum tensioner span tension.

An oblong sprocket.

The present invention includes a synchronous belt drive system, which includes: a first obround sprocket (10) having a toothed surface and at least one linear portion (16) disposed between two arcuate portions (14, 15), said arcuate portions having constant radii (R1, R2), said linear portion having a predetermined length; a sprocket (300) having a toothed surface, the sprocket being engaged to a first obround sprocket by an annular toothed member (200), and an amplitude and a phase of the first obround sprocket (10) being such that an angular displacement timing error between the sprocket and the first obround sprocket is less than 1.5 degrees peak-to-peak.

FIG. 1 is a side view of an oblong sprocket. The sprocket 10 of the present invention includes a toothed surface 11. The toothed surface 11 engages a toothed belt. The toothed surface 11 comprises a raised area 12 and an adjacent groove 13. The grooves 13 have a shape compatible with the toothing of a correspondingly designed toothed belt. Toothed belts are also known as synchronous belts because they are used to synchronize the rotation of the drive and driven sprockets.

Sprocket 10 includes a portion 14 and a portion 15. Portion 14 has an arcuate toothed surface 11a which includes a constant radius R2. Portion 15 has an arcuate toothed surface 11b which includes a constant radius R1. Since the radii R1 and R2 are equal and constant, the portions 14 and 15 are segments of a circle. The use of the circular segments in this manner reduces the complexity of the design and manufacturing process of the inventive sprocket.

A linear portion 16 is disposed between portions 14 and 15. The portion 16 comprises a rectangular section having the effect of displacing each portion 14 and 15 from each other, thereby imparting an oblong shape to the sprocket. Sprocket surface 11 is straight, i.e., linear or flat, between points 160 and 161 and between points 162 and 163.

The length of the flat portion 16 is related to the system torque ripple amplitude. In this embodiment, portion 16 has a dimension (W) of about 2mm between points 160 and 161 and between points 162 and 163. Thus, the center of curvature 17 of the portion 14 is displaced from the center of rotation 20 of the sprocket by a distance of W/2 (about 1mm), which is referred to as the "amplitude" of the eccentricity of the obround sprocket. Furthermore, the center of curvature 18 of the portion 15 is displaced from the center of rotation 20 of the sprocket by a distance of W/2 (about 1 mm). The dimensions given are for illustrative purposes only and are not intended to be limiting. It also follows that the long axis of the sprocket has a length dimension:

L_{long and long}＝R1+R2+W。

The preferred section (MG) of each portion 14, 15 has the dimensions:

MG ═ R1+ W/2 or (R2+ W/2).

The short axis of the sprocket has a length dimension:

L_{short length}＝R1+R2

The more general definition of the amplitude (which may also be more useful for sprockets having more than two lobes) is the difference between the largest good segment (major segment) and the smallest bad segment (minor segment) (i.e., MG-R1 or MG-R2 in the case of such double salient angles). In the case of a symmetrical lenticular angle, the amplitude is only W/2. In the case of an asymmetric design or with more than two lobes, there will be a small deviation from W/2.

The length (W) of portion 16 is determined by the radii of portions 14 and 15 and depends on the dynamic angular vibration characteristics being cancelled, which are described elsewhere in this specification. The sprocket 10 can be designed using a constant surface pitch, a constant angular pitch, or a combination of the two. "surface pitch" is defined as the distance measured about an outside diameter line between any two consecutive, corresponding "nodes" on the outside diameter of the sprocket. The constant surface pitch is calculated as follows:

SP ═ ((((Ng × nominal pitch)/Pi) -PLD) × Pi)/Ng)

Wherein

SP is the pitch of the surface,

ng is the number of grooves in the sprocket,

nominal pitch is the nominal system pitch,

pi is-3.141, and

PLD (radial PLD) of system

An "angular pitch" is defined as the angular difference between any two consecutive, corresponding "nodes" on a sprocket, and can be measured in degrees or radians.

The constant angular pitch is defined as follows:

AP 360/Ng (degrees),

wherein

AP is the angular pitch, and

ng is the number of grooves in the sprocket

The sprocket groove profiles can be individually designed to suit the specific dynamics of the engine.

The modulus of elasticity of the span of the belt is optimized in combination with the tooth modulus and the sprocket offset (W/2) to substantially reduce or eliminate tension fluctuations at a predetermined engine speed. Thus, in this application, the belt is analyzed and designed as a spring member of the system, in addition to being sized to transmit the required tensile load. The system dynamic response is selected by an iterative process to arrive at a combination of belt modulus and oblong sprocket radius (R1 and R2) that substantially reduces or eliminates all tension fluctuations that would otherwise be transmitted through the belt and belt drive system.

FIG. 2 is a side view of an alternate embodiment of the sprocket. This embodiment comprises three linear sections arranged between arcuate portions 14, 15, 16 as otherwise depicted in fig. 1. Three linear sections (161 to 162) and (163 to 164) and (165 to 166) are disposed between each arcuate portion 14, 15, 16. Each arcuate portion 14, 15, 16 includes a constant and equal radius R1, R2, R3, respectively. The three linear sections are evenly spaced about the circumference of the sprocket at approximately 120 ° intervals. Fig. 9 represents the 1.5 order load characteristic in a system using the sprocket shown in fig. 2.

Fig. 3, 4 and 5 are some typical drive layouts for a four-cylinder, four-stroke internal combustion engine using a toothed belt system to drive the camshaft and auxiliary equipment. These engines typically have high 2 nd order dynamics. Some diesel engines may have a dominant 1.5 step depending on fuel pump specifications. A schematic diagram illustrating such dynamics can be seen in fig. 7, 8 and 9.

To counteract the 2 nd order dynamics, the inventive sprocket 10 is attached to the engine crankshaft CRK. Depending on the presence of other dominant steps, it may be desirable to implement alternative embodiments of the sprocket. These sprockets may be attached to the crankshaft, but could equally be applied elsewhere in the system, such as on the water pump, on the fuel pump, or on the camshaft sprocket(s). The engine crankshaft is the drive for the entire belt drive system. The drive direction of the belt is DoR. Due to the sprocket gear ratio, the engine crankshaft CRK rotates twice for each rotation of the camshaft CAM 1.

In FIG. 3, sprocket 300 is connected to camshaft CAM1 and sprocket 304 is connected to second camshaft CAM 2. Idlers IDR1 and IDR2, as known in the art, are used to maintain proper belt lay-out and tension control. The sprocket 100 is connected to the water pump WP. The belt 200 is wrapped between a plurality of sprockets. The direction of rotation of the belt 200 is shown as DoR. The point at which belt 200 engages crankshaft sprocket CRK is 201. Camshaft inertia and torque load are represented by 301.

The toothed belt 200 is wrapped between the sprocket 10 and the cam sprocket 300. The belt entry point 201 is the point where the belt 200 engages a sprocket. The belt span length between the crankshaft CRK and the cam sprocket 304 is "SL".

Similarly, in fig. 4 and 5, camshaft sprocket 300 is attached to engine camshaft CAM. In fig. 4, the load characteristic 301 contains the torque characteristic of the fuel pump attached to the rear portion of the camshaft, and in fig. 5, the fuel pump torque is represented by a load characteristic 302. There may also be inertial and torque loads (301, 302, 101) caused by other components, such as water and vacuum pumps (i.e., WP (101) in fig. 4 and 5). In fig. 4, IDR1 and IDR2 are idler pulleys known in the art to suitably guide belt 200. In fig. 4, the belt span length between the crankshaft sprocket 10 and the cam sprocket 300 is "SL".

For gasoline engines, the prevailing cyclic fluctuating torque load is typically characteristic of the camshaft. For diesel engines, the dominant step may be produced by the camshaft and/or a fuel injection pump that may be included in the drive system. The torques induced by the water pump and vacuum pump may vary, but these torques are not themselves periodic at the same period or frequency as the camshaft and are generally not the primary characteristic of drive dynamics.

Fig. 5 is a perspective view of an additional single cam engine embodiment having a fuel injection pump (incorporated into the transmission of a diesel engine). In this embodiment, in addition to the system shown in fig. 4, the system further includes a sprocket 305 connected to the fuel pump IP. Also shown is a sprocket P1 that can be engaged with another multi-ribbed belt used to drive various engine accessories (not shown). In fig. 5, the cam load is denoted by 301, and the fuel pump load is denoted by 302. The sprocket 100 is connected to the water pump WP. In fig. 5, the torque load caused by the fuel injection pump is represented by 302.

In fig. 7, a typical total load characteristic of a four-cylinder, four-stroke engine is represented by curve "E". Curves "D" and "C" represent typical 2 nd and 1.5 th order characteristics extracted from the total load characteristics. The load characteristics of an in-line four cylinder, four stroke gasoline powered engine typically do not include 1.5 steps.

The change in the average radius at the belt engagement point 201 of the inventive sprocket 10 as it rotates is curve "C" in fig. 8 and 9. The integral of curve "C" (which is the effective length change of the band in fig. 4) is curve "D" in fig. 8 and 9. The derivative of the change in average sprocket radius is the acceleration of a given point on the tooth surface 11 due to the change in sprocket shape.

To counteract the 2 nd order dynamics, the flat portion 16 of the obround sprocket 10 is arranged in a timed relationship with respect to the camshaft sprocket 300 such that the effective length of the belt 200 between sprocket 300 and sprocket 10 in fig. 4 is varied in a manner that substantially eliminates the alternating belt tension caused by the periodic camshaft torque fluctuations. As an example of a design that eliminates 2 nd order dynamics, this may be accomplished by timing the maximum length (R1+ R2+ W) of the sprocket 10 to coincide with the belt entry point 201 when camshaft torque is at a maximum (and therefore belt tension is at a maximum).

The absolute dimensional characteristics of a drive incorporating an obround sprocket depend on parameters such as fluctuating torque, belt span modulus, inertia of each driven accessory in the system, belt installation tension, and belt-sprocket interaction. The interaction between the belt and the sprocket depends on parameters such as the number of teeth meshing on the sprocket, the modulus of the belt teeth, the size of the belt, and the coefficient of friction between the belt and the sprocket surface.

FIG. 6 is a schematic diagram of a dual cam, four cylinder, four stroke gasoline engine. The illustrative system includes cams CM1, CM2, and a strap B wrapped between the cams. It also includes tensioner TEN, water pump WP and crankshaft sprocket CRK. The direction of rotation of band B is DoR. The span lengths of interest are between sprockets CRK and IDR, between sprockets IDR and sprockets WP, and between sprockets CRK and sprockets WP. In fig. 6, the belt span length between the crankshaft sprocket CRK and the cam sprocket CM1 is "SL". For computational purposes, they may be considered a span "SL" because there is no major load impact between CM1 and CRK along the DoR. Approximate typical values of the variables for the system depicted in fig. 6 are as follows:

typical cam torque fluctuations are: +40N/-30N

Belt span modulus: 240Mpa

Typical component inertia values are:

CRK＝0.4gm²

CM1＝CM2＝1.02gm²

WP＝0.15gm²

belt installation tension: 400N (the installation tension is preferably maintained by tensioner TEN using the tensioners described herein).

Meshing teeth on three sprockets:

a tooth;

and (4) teeth.

Size of the tape: the width is 25.4 mm; length 1257.3mm

Typical values for the coefficient of friction for the sprocket surface 11 are in the range of 0.15 to 0.5, typically 0.2.

Typical belt installation tension values may be in the range of 75N to 900N, depending on system requirements.

The belt span modulus depends on the configuration of the tensile members, the number of strands of the tensile members within the belt, and the belt width. For a 25.4mm wide ribbon having 20 fiberglass tensile members, an example of the ribbon span modulus is in the range of about 240 Mpa.

Fig. 7 represents typical total load characteristics of a driven sprocket of a four-cylinder, four-stroke diesel engine, which includes extracted 1.5-order curve (curve "C") and 2-order curve (curve "D"). The load characteristics of an in-line four cylinder, four stroke gasoline powered engine typically do not include 1.5 steps. "offset" refers to W/2. "Total load" refers to line "E" of FIG. 7.

In fig. 7, line "a" is zero torque. Line "B" depicts the average torque in the belt drive system. Curve "C" is a 1.5 order torque characteristic extracted from the total load curve "E". Curve "D" is a 2 nd order torque characteristic extracted from the total load curve "E". Curve "E" is the total torque characteristic of the engine measured at the crankshaft CRK. The area under curve "E" represents the work done to turn the engine at a particular speed.

Fig. 8 represents the 2 nd order load characteristic (curve "B") of a drive sprocket for a four cylinder, four stroke engine, which includes the change in radius of the obround sprocket (curve "C") and the resulting change in belt span length (curve "D").

In fig. 8, line "a" is zero torque. Curve "B" is a 2 nd order torque characteristic extracted from the total load. Curve "C" is the change in effective crankshaft pulley radius caused by segment 16 in fig. 1 as the crankshaft pulley rotates through 360 degrees. Curve "D" is the integral of curve "C" and is the effective change in belt drive span length caused by the sprocket depicted in fig. 1.

Fig. 9 represents a 1.5 order load characteristic "B" for a drive sprocket (or other driven device that would cause a 1.5 order) for a four cylinder, four stroke diesel engine having a three piston fuel pump, which includes the variation in sprocket radius length (curve "C") and resulting belt span length variation (curve "D") for an alternate three lobe embodiment of an obround sprocket (fig. 2). The belt span length is, for example, the distance between the CAM sprocket CAM and the crankshaft sprocket CRK of FIG. 6.

In fig. 9, line "a" is zero torque. Curve "B" is a 1.5 order torque characteristic extracted from the total load. Curve "C" is the change in effective crankshaft pulley radius as the crankshaft pulley rotates through 360 degrees. Curve "D" is the integral of curve "C" and is the effective change in drive length caused by the alternate embodiment of the sprocket depicted in fig. 3.

The modulus of elasticity of the tensile members of the various belts used in the system of the present invention is shown in FIG. 10. Curves SS1 through SS6 are referred to as stress-strain curves for the various ribbons 200. Each curve represents the modulus for a different material used for the tensile cord in the belt. The belt body of the elastomeric HNBR is shown, but is not limited thereto. In addition to HNBR, other belt body materials may include EPDM, CR (chloroprene), and polyurethane, or a combination of two or more of the foregoing materials. The material comprises:

SS1 (fiberglass #1 tensile cord, HNBR body)

SS2 (fiberglass #2 tensile cord, HNBR body)

SS3 (fiberglass #3 tensile cord, HNBR body)

SS4 (carbon fiber tensile cord, HNBR body)

SS5 (aramid fiber tensile cord, HNBR main body)

SS6 (carbon fiber tensile cord, HNBR body).

As is known in the art, the modulus of elasticity M of each tensile member is the slope of each curve (SS 1-SS 6). Typically, the measurement and calculation is performed on a substantially linear portion of the curve. In addition to glass fibers, carbon fibers, and aramid, additional tensile member materials may include fine stainless steel wire or PBO.

Delta stress/delta strain (measured in the approximately linear portion of the curve)

The belt span modulus depends on the configuration of the tensile members, the number of strands of the tensile members within the belt, and the belt width. An example of a belt span modulus for curve SS1 would be about 242Mpa for a 25.4mm wide belt with 20 strands of fiberglass tensile members.

Fig. 11 is a series of curves showing the effect of phasing/phase shift of the major length (major length) of an obround sprocket on the engine dynamics of the system in fig. 6. Curve "D" is the optimal timing arrangement between the location of the sprocket optimal length to the belt entry point 201 and the torque pulse. Curves A, B and C are the clockwise timing offsets of +6, +4, and +2 teeth, respectively, from the curve "D" position. Curve "E" is a timing shift of 2 teeth in the counterclockwise direction. The phasing of the maximum belt span length relative to peak torque and inertial load may vary depending on the dominant step of the drive and those steps the system is to reduce. The belt entry point 201 is the point at which the belt engages a sprocket. In fig. 3, the span length is "SL".

With respect to angular spacing or phasing, the allowable angular tolerance is calculated using:

+/- (360/2X number of sprocket grooves).

The belt drive span length is greatest when the torque is greatest.

Fig. 12 is a graph showing the effect of a correctly phased obround sprocket on a two-cam, four-cylinder, four-stroke engine such as that shown in fig. 6. Curves "a" and "B" represent measured values of angular vibration at the intake camshaft sprocket and at the exhaust camshaft sprocket, respectively, for a prior art design using a circular sprocket.

For comparison, curves "C" and "D" represent measured values of angular vibration at the intake camshaft sprocket and at the exhaust camshaft sprocket, respectively, in the case where the sprocket of the present invention is used on a crankshaft. The resulting reduction in angular vibration is about 50%.

Similarly, fig. 13 is a graph showing the effect of a correctly phased obround sprocket as depicted in fig. 1 on a two-cam, four-cylinder, four-stroke engine as shown in fig. 6. Curves "a", "B" and "C" represent measured values of maximum, average and minimum dynamic tight side tensions, respectively, over a range of engine speeds for a prior art drive design. In this example, the tension is measured at position IDR in fig. 6. To extend the useful life of the belt, the tight side tension of the belt should be minimized. Curves "D", "E" and "F" represent measured values of maximum, average and minimum belt tight side tensions with the sprocket of the present invention. The resulting reduction in installed tight side tension is in the range of 50-60% over the engine's resonant speed range (about 4000rpm to about 4800 rpm). The reduction in tight side tension of the belt has the potential to significantly increase the operational life of the belt.

The system of the present invention is useful in reducing timing errors in internal combustion engines. Timing error is a positional difference between the driving shaft and the driven shaft caused by random factors such as vibration, component inaccuracy, and elastic deformation. In this case, rotational inaccuracy of the camshaft (driven) of the internal combustion engine compared to the crankshaft (driven) of the engine is caused. Typically recorded in units of peak-to-peak degrees. For example, referring to fig. 3, sprocket 300 and sprocket 304 are each oblong. The use of an oblong sprocket significantly reduces timing errors, which in turn improves fuel economy, reduces emissions, and generally improves engine performance and efficiency. At the component level, the reduced timing error and reduced system load result in better durability and a lower likelihood of NVH problems. The reduction in tension reduces NVH levels in the drive, particularly the engagement sequence. The application of an oblong sprocket to reduce timing error is not limited to camshafts for engines. Benefits may be equally achieved by inserting an oblong sprocket on the crankshaft or fuel pump.

FIG. 14 is a graph of angular vibration versus crankshaft speed. The exemplary angular vibration decreases as engine speed increases. Fig. 14 shows data for a motor driven (motored) engine test rig and an ignited engine test rig. For a motor-driven engine, the crankshaft is driven by the motor and there is no combustion of fuel in each cylinder. For an ignited engine, the crankshaft is driven in the conventional manner for an internal combustion engine (i.e., with combustion of fuel in each cylinder). For a given engine speed, the angular vibration reflected by the motor-driven engine equipment (MER) is less than the angular vibration of the ignited engine equipment (FER).

Fig. 15 is a graph of vibration angle versus crankshaft speed for an intake cam. An oblong sprocket is mounted to the intake valvetrain camshaft. Three states are shown. The first is for a standard drive system without an obround sprocket (curve a). The second had an oblong sprocket (curve B) and the third had an oblong sprocket and a high modulus belt (curve C). The phase and amplitude of the oblong sprocket was 10.5 pitches and 1.5mm from the 3 o' clock position. The modulus of the standard tape was 630,000N and the modulus of the high modulus tape was 902,000N. Modulus is given in newtons (N) and is defined as the force required to extend a unit length by 100%.

The vibration angle of the third state (curve C) is significantly reduced to less than 0.5 degrees peak-to-peak, both measured at 4000RPM, when compared to the value of a standard driver system at about 1.5 degrees peak-to-peak.

Fig. 16 is a graph of vibration angle versus crankshaft speed for an exhaust cam. The oblong sprocket is mounted to an exhaust valve train camshaft. Three states are shown. The first is for a standard drive system without an obround sprocket (curve a). The second had an oblong sprocket (curve B) and the third had an oblong sprocket and a high modulus belt. The vibration angle of the third state (curve C) is significantly reduced to about 0.5 degrees peak-to-peak, both measured at 4000RPM, when compared to the value of a standard driver system at about 1.5 degrees peak-to-peak. However, depending on the engine, the improvement may range from less than 1.5 degrees peak to about 0.5 degrees, with a reduction of slightly more than 60%. The phase and amplitude of the oblong sprocket was 23.5 pitches and 1.5mm from the 3 o' clock position. The modulus of the standard tape is about 630,000N and the modulus of the high modulus tape is about 902,000N.

FIG. 17 is a graph of angular displacement versus crankshaft speed for an intake cam. Angular displacement is also known as timing error and is measured relative to crankshaft position. An oblong sprocket is mounted to the intake valvetrain camshaft. Three states are shown. The first is for a standard drive system without an obround sprocket (curve a). The second had an oblong sprocket (curve B) and the third had an oblong sprocket and a high modulus belt (curve C). The angular displacement of the third state (curve C) is significantly reduced to less than 0.5 degrees peak-to-peak, both measured at 4000RPM, when compared to the value of a standard drive system at about 1.5 degrees peak-to-peak. However, depending on the engine, the improvement may range from less than 1.5 degrees peak to about 0.5 degrees, with a reduction of slightly more than 60%. The phase and amplitude of the oblong sprocket was 10.5 pitches and 1.5mm from the 3 o' clock position. The modulus of the standard tape is about 630,000N and the modulus of the high modulus tape is about 902,000N.

FIG. 18 is a graph of angular displacement versus crankshaft speed for an exhaust cam. The oblong sprocket is mounted to an exhaust valve train camshaft. Three states are shown. The first is for a standard drive system without an obround sprocket (curve a). The second had an oblong sprocket (curve B) and the third had an oblong sprocket and a high modulus belt (curve C). The angular displacement of the third state is significantly reduced to about 0.5 degrees peak-to-peak, both measured at 4000RPM, when compared to the value of a standard drive system at about 1.5 degrees peak-to-peak. However, depending on the engine, the improvement may range from less than 1.5 degrees peak to about 0.5 degrees, with a reduction of slightly more than 60%. The phase and amplitude of the oblong sprocket was 23.5 pitches and 1.5mm from the 3 o' clock position. The modulus of the standard tape is about 630,000N and the modulus of the high modulus tape is about 902,000N.

FIG. 19 is a graph illustrating the effect of oblong sprocket phasing on various camshaft timing errors. The Y-axis is the angular displacement or timing error of each cam sprocket relative to the crankshaft. It is referred to as the peak-to-peak value, i.e. the difference in value between the minimum and maximum values. Columns 1 and 2 of the drawings record standard drive set-ups all using circular sprockets. Column 3 records use 3 rd order obround sprockets mounted on the intake and exhaust camshafts. Each sprocket is positioned such that the maximum offset corresponds to the camshaft lobe. Columns 4 to 13 record various tests using oblong sprockets of different offsets. The "3 o' clock" position is the reference point for all offsets. The values given are simply the pitch or number of grooves through which the sprocket reference point rotates from that position plus "g". A "reference point" is a point that is used as a reference for angle measurement. This is set at the 3 o' clock position. "CW" refers to the clockwise direction. For example, "EX 23.5g CW" refers to the 3 o 'clock position and the exhaust cam obround sprocket is offset 23.5 grooves in a clockwise direction from the 3 o' clock position on the engine.

FIG. 20 is a graph showing the effect of oblong sprocket phasing on individual camshaft timing errors with a standard high modulus belt. The Y-axis is the angular displacement (in degrees peak to peak) or timing error of each cam sprocket relative to the crankshaft. It is referred to as the peak-to-peak value, i.e. the difference in value between the minimum and maximum values. Columns 1 and 3 of the drawings record standard drive set-up using all round sprockets. Each column records the use of 3-step oblong sprockets mounted on the intake and exhaust camshafts. Each sprocket is positioned such that the maximum offset corresponds to a camshaft lobe. Columns 2 and 4 to 8 record various tests using oblong sprockets of different offsets. The "3 o' clock" position is the reference point for all offsets. The values given are simply the pitch or number of grooves through which the sprocket reference point rotates from that position. A "reference point" is a point that is used as a reference for angle measurement. Which is set at the 3 o' clock position. The phase of the oblong sprocket for exhaust was 23.5 pitches from the 3 o 'clock position, the phase of the oblong sprocket for intake was 10.5 pitches from the 3 o' clock position, and the amplitudes of the oblong sprockets for exhaust and intake were each 1.5 mm. The modulus of the standard tape is about 630,000N and the modulus of the high modulus tape is about 902,000N.

Fig. 21 is a graph showing the effect of an oblong sprocket on timing error due to belt width. Column 1 records a 14mm wide belt in a system using a circular sprocket. Column 2 records a 14mm wide belt in a system using an oblong sprocket. Column 3 records a 14mm wide belt using high modulus belts in a system using standard sprockets. Column 4 records a 14mm wide belt using a high modulus belt in a system using an oblong sprocket. Column 5 records an 18mm wide belt using standard modulus belts in a system using standard sprockets.

FIG. 22 is a graph illustrating the effect of the magnitude of the oblong sprocket on timing error due to eccentricity. Each column records an oblong sprocket used on the intake camshaft and the exhaust camshaft. The magnitude of eccentricity for each system is between 1.0mm and 1.5 mm.

Tests can be performed on both motor-driven and ignited engines to verify the effectiveness of the obround sprocket in reducing belt drive system dynamics. The improved results for timing error contained in the figures are generated on a motor-driven engine. Although in most cases these results are shifted to ignited engines, in some cases, the oblong sprocket does not reduce the dynamics of some engines. Tests should be conducted on the fired engine to ensure that the required improvements are achievable and reliable. The steps required to perform the test are known in the field of engine dynamics. These also include the need for the vibration sensor to operate in an oil environment, to be able to withstand temperatures up to 160 ℃, and to withstand chemical attack from oils and additives. A consistency check is performed at the beginning and end of each series of test runs. Measurements were taken during start-up with 60 seconds of acceleration from idle to maximum engine speed. Data capture and analysis can be performed using a standard Rotec system.

A tensioner.

The invention comprises a tensioner comprising: a base having an axially extending cylindrical portion including a radially outer surface and a receiving portion radially inward of the radially outer surface; an eccentric arm pivotally engaged with the radially outer surface; a torsion spring disposed within the radially inner receiving portion, the torsion spring applying a biasing force to the eccentric arm; and a pulley having a shaft diameter connected to the eccentric arm.

Fig. 23 is an exploded view of the preferred tensioner. Tensioner 400 includes a base 410. The base 410 includes an axially extending cylindrical portion 412 having an outer surface 414. Cylindrical portion 412 further includes an opening 411 and a receiving portion 418.

The eccentric arm 420 pivots about the cylindrical portion 412. A bushing 460 is disposed between the inner surface 422 and the outer surface 414. Bushing 460 includes a slot 461 that is generally aligned with opening 411 in cylindrical portion 412. Pulley 440 is journaled to surface 421 on needle bearing 450. Needle bearings are used in an oil bath environment. Other bearings known in the art are also suitable.

A torsion spring 430 engages the over-center arm 420 and biases the over-center arm toward the belt (not shown) to apply a belt load. The end 431 protrudes through the slot 461 and the opening 411 to engage the receiving portion 424 of the eccentric arm 420. The end 432 engages the end receiving portion 415 in the base 410. The torsion spring 430 is disposed entirely within the receiving portion 418. The receiving portion 418 is a central hollow portion of the cylindrical portion 412. Torsion spring 430 is coplanar with bearing 450, pulley 440, and eccentric arm 420. The torsion spring 430 is disposed radially inward of the pulley 440, the bearing 450, the bushing 460, and the cylindrical portion 412. That is, the torsion spring 430, the bearing 450, the pulley 440, and the eccentric arm 420 are all concentrically arranged such that none of the listed components are axially displaced from each other along the axis a-a.

The retaining ring 406 engages a circumferential slot 416 in the base 410. The retaining ring 405 engages a circumferential slot 423 in the eccentric arm 420. Retaining ring 405 retains bearing 450 to eccentric arm 420. The retaining ring 406 retains the eccentric arm 420 to the base 410. In the presence of oil, the retaining rings 405 and 406 may each act as a thrust washer to transfer axial forces.

Pulley 440 is press fit onto bearing 450. Fastener 404 protrudes through torsion spring 430 and hole 417 in base 410 to secure tensioner 400 to a mounting surface, such as an engine (not shown).

The bushing 460 includes a coefficient of dynamic friction (COF) in a range of about 0.05 to about 0.20. The static COF is preferably lower than the dynamic COF.

Fig. 24 is an exploded top view. Eccentric arm 420 pivots about an axis a-a that is centered on cylindrical portion 412 and protrudes through fastener 404. The eccentric arm 420 pivots about axis a-a. Pulley 440 rotates about "B," which is the geometric center of eccentric arm 420. "B" is eccentrically offset from axis a-a, allowing eccentric pivoting movement of eccentric arm 420, which in turn allows tensioner 400 to apply a variable load to a belt (not shown).

Fig. 25 is a perspective view of the base. The end receiving portion 415 is provided at one end of the receiving portion 418 in the base 410. End 432 engages end receiving portion 415, thereby securing end 432 and acting as a reaction point for the torsion spring.

Fig. 26 is a perspective view of an eccentric arm. "B" is the geometric center of the eccentric arm 420 and is the point about which the pulley 440 rotates. The eccentric arm 420 pivots about an "a" on axis a-a. The receiving portion 424 of the over-center arm engages the end 431 of the spring 430.

Fig. 27 is a perspective view of a torsion spring. The end 431 protrudes into the receiving portion 424 of the eccentric arm 420. End 432 engages end receiving portion 415.

Fig. 28 is a cross-sectional view of the tensioner. The torsion spring 430, bushing 460, cylindrical portion 412, eccentric arm 420, bearing 450, and pulley 440 are all concentrically arranged such that none of the listed components are axially displaced from one another along axis a-a. This completely concentric and nested arrangement minimizes the height (or width) of the tensioner, allowing the tensioner to be used in very narrow applications.

Fig. 29 is an exploded view of an alternative embodiment. The components are the same as described herein, except that bearing 451 is a sliding bearing and bushing 460 is omitted. The alternative embodiment is configured to operate in oil and/or be used with oil splash lubrication. The eccentric arm 420 pivots about axis a-a. The pulley 440 rotates about an axis B-B, see fig. 26. Axis a-a is disposed away from axis B-B and, therefore, axis B-B is not coaxial with axis a-a, thereby allowing eccentric pivoting movement of eccentric arm 420.

Fig. 30 is a top view of the alternative embodiment of fig. 29.

Fig. 31 is a cross-sectional view of the alternative embodiment of fig. 29. Torsion spring 430, eccentric arm 420, and bearing 451 are concentrically arranged such that none of the listed components are axially displaced from each other along axis a-a. Fluid conduit 471 in base 410 provides a path for fluid, such as oil, to flow from an engine oil system (not shown) to bearing 451 via fluid conduit 473 to lubricate the bearing. O-ring 472 provides a means to seal the connection to the engine oil system.

Fig. 32 is a side view of an alternative embodiment. Instead of eccentric arm 420 and pulley 440, this alternative embodiment includes cam 445. The cam 445 operates on the same principle as the eccentric arm 420 and occupies the same position in the device. Pulley 440 is not present. The cam 445 engages the elongated member 480. The elongate member 480 may comprise any suitable low friction material known in the art. The elongated member 480 may also be referred to as a sliding guide. The chain "C" slidingly engages a surface of the slide guide 480. The pivot 481 is provided at one end of the slide guide. The slide guide 480 pivots about the pivot 481 in response to rotation of the cam 445. Due to the eccentric form of the surface 446, rotation of the cam 445 causes the slide guide 480 to pivot about the pivot 481, thereby maintaining the load on the chain "C". For example, this embodiment is useful in an internal combustion engine timing system. This embodiment may be used with a timing belt instead of a chain.

Fig. 33 is a perspective view of the alternative embodiment of fig. 32. A surface 446 of the cam 445 engages the slide guide 480.

And a synchronous belt driving system.

The drive system of the present invention comprises a belt, at least one obround sprocket as described above, and a tensioner as described above. The drive system comprises at least two sprockets, a drive sprocket and a driven sprocket. As described herein, at least one of the drive sprocket and the driven sprocket is oblong. The tensioner is preferably designed as described herein and may utilize a backside idler or slide to contact the belt span.

In one embodiment, the synchronous belt drive system is an overhead cam drive for an internal combustion engine, such as for an automobile or other land vehicle. Examples are introduced above and shown in fig. 3-6. Fig. 3 and 6 represent an exemplary dual overhead cam drive system, while fig. 4 and 5 represent a single overhead cam drive system. Each of these examples has a crank sprocket as the driver and one or more cam sprockets as the driven sprockets. Each drive may also contain various back side idler devices and water pumps as well as tensioners as described herein. The actuator of fig. 5 comprises a jet pump. The driver may be used with a diesel or gasoline engine. In other embodiments, the drive may be for a single driven component, such as a pump or balance shaft, which contains, for example, a crank sprocket, a driven sprocket for the driven component, and a tensioner. In each case, as well as in other countless variations of drive systems, the inventive principles described herein may be used to provide a synchronous drive system that may operate dry or in fluid contact with oil or other engines, and may have a significantly narrower packaging width than previous drives.

The primary advantage of the above tensioner design for a synchronous belt drive system is compactness. In conventional belt drives, the tensioner is typically the widest component, thus significantly limiting the overall package width. With the tensioner design described herein, the belt can be made as small as possible, and the tensioner and other sprockets can actually be the same width plus a small amount of clearance, or can be slightly wider if desired to accommodate some run-out or potential misalignment.

The use of one or more oblong sprockets in combination with a high modulus timing belt forms such a drive: the drive is able to operate surprisingly well with narrow width belts. In the art, a reduction in belt width is typically associated with an increase in tension and load (per unit width) on the belt, resulting in increased span length deflection and increased tooth deflection, which in turn results in poorer timing accuracy and poorer durability. However, the present system may exhibit greatly improved timing associated with reduced tension and load on the belt, as well as adequate durability.

Examples are given.

For each of the following examples, the test layout was a dual overhead cam drive on a three cylinder engine driven by an electric motor coupled to the crankshaft, as shown in fig. 35. The three cylinder engine is known as a 1.0 liter FOX engine manufactured by ford motor company. The stock oil wet timing drive system was modified to accommodate the various test drives described below. All test drives were run on oil wetlands with Castrol Magnatec 5W-20 oil at an oil temperature of 140 ℃. The test layout 240 contained a 19-slot drive or crank pulley 243 and two 38-slot driven cam pulleys 242, 245 having an RPX slot profile and a 9.525-mm pitch to match the belt 200 with an RPP tooth profile. The speed profile and belt width, pulley and tensioner 244 details vary as described below.

Three exemplary strips shown in table 1 were used in each system test. Belt 1 is a stock belt supplied by Dayco for a 1.0 liter FOX engine. Belt 2 is an oil resistant belt supplied by Gates Unitta Asia. The belt 3 is a modified version of the belt 2 in which the tensile cords are replaced by hybrid cords made of carbon fibers wrapped with U-glass fibers (high strength glass).

The belt 1 is believed to be constructed as described by Dayco in WO 2005080820.

Belt 2 is believed to be constructed using belt materials as described in U.S. patent publication No. 2014/0080647A1 to Yamada et al, which is incorporated herein by reference. The facing fabric was a 2x2 twill with a para-aramid/nylon elastomeric yarn in the weft and nylon woven in the warp. The facing fabric is treated with an epoxy + NBR latex + hardener treatment as described in U.S. patent publication No. 2014/0080647a1 to Yamada et al. The treatment was applied by dipping and dried in a conventional oven. The back fabric was a woven 2x2 twill nylon 66 elastic fabric treated with NBR latex based RFL. The rubber composition of the belt body (teeth and back) is based on a nitrile group-containing copolymer rubber, i.e. HNBR comprising a short fiber reinforcement, resorcinol and a melamine compound. The rubber composition of the belt body further comprises additional ingredients known in the art including carbon black, certain plasticizers, antidegradants, curing agents and auxiliaries. The tensile cord 18 for the belt of system a is a twisted high strength glass fiber yarn treated with NBR RFL treatment to have oil resistance for use in oil wet environments.

The construction of belt 3 is the same as belt 2, except for the different tensile cords. The glass cords of the belt 3 are replaced by hybrid carbon/glass cords as described in us patent No. 7,682,274, i.e. the belt has a carbon fiber core yarn surrounded by a plurality of high strength glass fiber yarns.

TABLE 1

First series of tests of the comparative system.

The test layout for this first series of tests contained a stock tensioner and a pulley inertia matched to the stock pulley. The cam pulley is circular and cut with a Diameter Pitch Line Difference (DPLD) of 1.5-mm, and the crank pulley has a DPLD of 1.45-mm. The cam sprockets are inertia matched to a stock pulley containing a stock VVT device. Tensioner 244 is a stock compact tensioner with a single eccentric center and symmetric damping that provides an installation tension of about 500N and has a smooth 62-mm diameter pulley. The tensioner pulley surface width is 24.5mm and the overall width of the tensioner is about 36 mm. This is the widest component, so the total package width of the contrast drive is about 36 mm. The crank speed was varied between 5000rpm and 6000rpm, alternately accelerating for 10 seconds and decelerating for 10 seconds.

As shown in table 2, five drive systems were tested, and the results of the sixth system were predicted based on the other five drive systems. Only the bands differ in this series. At three different widths: belts 1 were used at 18mm, 12mm and 10 mm. At two different widths: the tape 2 was used at 18mm and 10mm and results for a width of 12mm were predicted. Two similar test devices were used. For comparative systems 1 and 2, the system was run for a fixed time, and for the other systems, the system was run until a belt failure occurred. The results of systems 3 and 6 show that the run time of device 1 is generally longer than that of device 2. Such changes in system testing are not uncommon. These tests are used as a benchmark for testing of the system of the present invention. Further testing was conducted with 10-mm wide tape due to the goal of minimizing system package size and the desire to avoid lengthy run times.

TABLE 2

¹See table 1.

²The test was suspended without tape failure.

³And predicting the service life.

⁴For systems 3, 5 and 6, the belt ran to full failure and therefore no tensile test could be performed.

A second series of tests of the comparative system versus the system of the present invention.

The test layout for this second series of tests is the same as the first series, i.e., as shown in fig. 35, whereas the test device is the third device, i.e., device 3. Device 3 differs from devices 1 and 2 in that a stock flywheel is included on device 3, but not on devices 1 and 2. Two drive systems were tested on the device 3 as shown in table 3. The comparison system 7 comprises a stock tensioner and a pulley inertia matched to the stock pulley. The cam pulley is circular and cut with a Diameter Pitch Line Difference (DPLD) of 1.5-mm, and the crank pulley has a DPLD of 1.45-mm. The cam sprockets are inertia matched to a stock pulley containing a stock VVT device. Tensioner 244 is a stock compact tensioner with a single eccentric center and symmetric damping that provides an installation tension of about 500N and has a smooth 62-mm diameter pulley. The tensioner pulley surface width is 24.5mm and the overall width of the tensioner is about 36 mm. This is the widest component, so the total package width of the contrast drive is about 36 mm. For this series of tests, the crank speed was held constant at 4750rpm (which is about the maximum resonant speed) during each test. The belt 2 was used in a comparison system 7 with a width of 10-mm, an RPP profile of 116 teeth and a pitch of 9.525 mm.

The inventive system 8 has the same test layout, however the cam sprocket is oblong and cut with a Diameter Pitch Line Difference (DPLD) of 1.5-mm. As shown in fig. 2, the oblong intake cam sprocket 242 is of a tri-lobe type with an eccentricity of 0.75-mm in magnitude and 13 gcw in phase (i.e., an angular offset of 13 slots in a clockwise direction from a reference position on the engine). As shown in fig. 2, the oblong exhaust cam sprocket 245 is of a tri-lobe type with an amplitude of 1.0-mm and a phase of 23.5g cw (i.e., an angular offset of 23.5 slots in a clockwise direction from a reference position). The cam sprockets are all inertia matched to the standing pulley.

The tensioner 244 of system 8 is a compact tensioner according to the design described herein, with a single eccentric center, providing an installation tension of about 500N and having a smooth 60mm diameter pulley. The tensioner therefore has a pulley surface width of 14mm, and an overall width of about 16 mm. The tensioner also has an asymmetric damping of about 300N, as described, for example, in U.S. Pat. No. 6,609,988B1 to Liu et al.

The crank pulleys are identical, i.e. have a reduced surface width to fit the narrow belt and package width.

The belt 3 was used in a system 8 with a width of 10-mm, an RPP profile of 116 teeth and a pitch of 9.525 mm.

Both systems are run until a belt failure occurs. The results are shown in table 3. The 690 hour run time for comparative system 7 is consistent with the 777 hour results for comparative system 6. The result of the inventive system 8 for a run time of 1266 hours was almost double. This shows that the combination of a narrow high modulus belt with an oblong sprocket and a special narrow tensioner can significantly improve belt life.

Table 3 contains some timing error results, including before and after testing. The improved timing error results can be considered a combination of effects including the effect of higher belt modulus, the effect of tensioner design, and the effect of an oblong sprocket. Table 3 shows some of these individual effects from other tests, some of which were performed on 12-mm wide strips as shown in table 3. The net result is that very narrow (10-mm belt) systems have maximum timing errors well below 1 peak-to-peak over the life of the belt.

In addition, table 3 shows that the effective tension of the belt in the system is significantly reduced from 500N to about 250N, and that the effect is primarily due to the oblong sprocket. This advantageous effect of the oblong pulley is very significant. It enables the mounting tension to be reduced without the risk of tooth jump, which in turn will reduce the wear rate of the tooth tip.

Thus, when a high modulus belt is combined with an oblong sprocket and concentric tensioner, the results of systems 7 and 8 show superior overall system performance and belt life. It is believed that the results can be even further improved by optimizing the mounting tension. The oblong sprocket significantly reduces the effective tension at the crank, which will generally extend belt life against tooth shear failure modes. However, since the observed failure mode is tooth tip wear, which is more related to PV (contact pressure times slip speed) in the system, the positive effect of the oblong sprocket on belt life is reduced. In other words, if we use an obround sprocket to partially reduce the effective tension (i.e., reduce tooth load) but also reduce the installation tension (i.e., lower PV between belt and pulley), the belt life can be further increased.

A strip width of 10-mm was chosen to speed up the test. From these results, and other experience regarding the correlation between accelerated testing and more practical applications, it is believed that increasing the strip width to about 14mm will result in a system 8 on the test apparatus with a durability of up to about 3500 hours, which is expected to correlate to a vehicle life of about 240,000 km. A belt width of 14-mm can be accommodated on the tensioner and pulley as described above, so that the total package width of the drive is about 18mm or less, which is about half the package width of a conventional drive with stock components. It is also estimated that the drive system can be easily manufactured to be about 30% lighter than a conventional drive system with stock components.

TABLE 3

¹See table 1.

²Separate tests were performed using 12-mm wide tapes.

Some additional aspects of the invention relating to the obround sprocket(s) are specifically set forth below.

Aspect 1. The present invention relates to a synchronous belt driving system, which comprises: a first oblong sprocket (10) having a toothed surface and at least one linear portion (16) disposed between two arcuate portions (14, 15), said arcuate portions having a constant radius (R1, R2), said linear portion having a predetermined length; a sprocket (300) having a toothed surface, the sprocket being engaged to the first obround sprocket by an endless toothed member (200); and the amplitude and phase of the first oblong sprocket (10) is such that the angular displacement timing error between the sprocket and the first oblong sprocket is less than 1.5 degrees peak-to-peak.

Aspect 2. The synchronous belt drive system as in aspect 1 further comprises: a second obround sprocket connected to a second rotational load, the second obround sprocket engaged with the endless toothed member; and the amplitude and phase of the second oblong sprocket is such that the angular displacement timing error between the sprocket and the second oblong sprocket is less than 1.5 degrees peak-to-peak.

Aspect 3. The synchronous belt drive system of aspect 1, wherein the angular displacement timing error between the sprocket and the first obround sprocket is less than 0.5 degrees peak-to-peak.

Aspect 4. The synchronous belt drive system of aspect 3, wherein the angular displacement timing error between the sprocket and the second obround sprocket is less than 0.5 degrees peak-to-peak.

Aspect 5. The synchronous belt drive system as in aspect 1, wherein a width of the endless tooth member is equal to or greater than 12 mm.

Aspect 6. The synchronous belt drive system of aspect 1, wherein the endless toothed member has a modulus in a range of about 630000N to about 902000N.

Aspect 7. The synchronous belt drive system of aspect 1, wherein the amplitude of the first obround sprocket is in the range of about 1.0mm to 1.5 mm.

Aspect 8. The synchronous belt drive system of aspect 1, wherein the phase of the first obround sprocket is in a range of 9 grooves to 25 grooves rotated with respect to a reference point.

Aspect 9. The synchronous belt drive system of aspect 8, wherein the reference point is referenced to a 3 o' clock position.

Aspect 10. The synchronous belt drive system of aspect 2, wherein the second obround sprocket is in phase in a range of 9 grooves to 25 grooves rotated with respect to a reference point.

Aspect 11. The synchronous belt drive system of aspect 10, wherein the reference point is referenced to a 3 o' clock position.

Aspect 12. The synchronous belt drive system of aspect 10, wherein the phase of the first obround sprocket is in a range of 9 grooves to 25 grooves rotated with respect to a reference point.

Aspect 13. The synchronous belt drive system of aspect 12, wherein the reference point is referenced to a 3 o' clock position.

Aspect 14. The synchronous belt drive system of aspect 1, wherein the sprocket is connected to a drive and the first obround sprocket is connected to a rotating load.

Aspect 15. The synchronous belt drive system of aspect 14, wherein the drive is an engine crankshaft.

Aspect 16. The synchronous belt drive system of aspect 2, wherein the first obround sprocket is coupled to an exhaust camshaft.

Aspect 17. The synchronous belt drive system of aspect 2, wherein the second obround sprocket is coupled to an intake camshaft.

Aspect 18. The present invention also relates to a synchronous belt drive system comprising: a first obround sprocket having a toothed surface and at least one linear portion disposed between two arcuate portions, the arcuate portions having a constant radius, the linear portions having a predetermined length; a sprocket having a toothed surface, the sprocket being engaged to the first oblong sprocket by an endless toothed member; the amplitude and phase of the first oblong sprocket are such that the angular displacement timing error between the sprocket and the first oblong sprocket is less than 1 degree peak-to-peak; a second obround sprocket connected to a second rotational load, the second obround sprocket engaged with the endless toothed member; and the amplitude and phase of the second oblong sprocket is such that the angular displacement timing error between the sprocket and the second oblong sprocket is less than 1.5 degrees peak-to-peak.

Aspect 19. The synchronous belt drive system of aspect 18, wherein the first oblong sprocket is connected to an exhaust camshaft, the second oblong sprocket is connected to an intake camshaft, and the sprockets are connected to an engine crankshaft.

Aspect 20. The synchronous belt drive system of aspect 19, wherein the angular displacement timing error between the sprocket and the first obround sprocket is less than 0.5 degrees peak-to-peak, and the angular displacement timing error between the sprocket and the second obround sprocket is less than 0.5 degrees peak-to-peak.

Some additional aspects of the invention relating to tensioners are specifically set out below.

Aspect 1. The present invention relates to a tensioner comprising: a base having an axially extending cylindrical portion including a radially outer surface and a receiving portion radially inward of the radially outer surface; an eccentric arm pivotally engaged with the radially outer surface; a torsion spring disposed within the radially inner receiving portion, the torsion spring applying a biasing force to the eccentric arm; and a pulley journalled to the eccentric arm.

Aspect 2. The tensioner of aspect 1, wherein the pulley is journaled on a needle bearing.

Aspect 3. The tensioner as in aspect 1, wherein the eccentric arm, the pulley, and the torsion spring are concentrically arranged such that none of the eccentric arm, the pulley, or the torsion spring is axially displaced from each other along axis a-a.

Aspect 4. The tensioner as in aspect 1, wherein the eccentric arm is journaled to the base on a bushing.

Aspect 5. The invention also relates to a tensioner comprising: a base cylindrical portion having a radially outer surface and a radially inner receiving portion; an eccentric arm pivotally engaged with the radially outer surface; a torsion spring disposed within the radially inner receiving portion, the torsion spring applying a biasing force to the eccentric arm; and an elongate member engaged with the eccentric arm and arranged to pivot in response to rotation of the eccentric arm.

Aspect 6. The tensioner as in aspect 5, wherein the eccentric arm and the torsion spring are concentrically arranged such that neither the eccentric arm nor the torsion spring is axially displaced from each other along axis a-a.

Aspect 7. The tensioner as in aspect 5, wherein the eccentric arm is journaled to the base on a bushing.

Aspect 8. The tensioner of aspect 5, wherein the pulley is journaled to the eccentric arm on a needle bearing.

Aspect 9. The invention also relates to a tensioner comprising: a base having an axially extending cylindrical portion including a radially outer surface and a radially inner receiving portion; an eccentric arm pivotally engaged with the radially outer surface; a torsion spring disposed within the radially inner receiving portion, the torsion spring applying a biasing force to the eccentric arm; a pulley having a shaft diameter connected to the eccentric arm; and wherein the eccentric arm, the pulley, and the torsion spring are concentrically arranged such that none of the eccentric arm, the pulley, or the torsion spring is axially displaced from any of the eccentric arm, the pulley, or the torsion spring along the axis a-a.

Aspect 10. The tensioner as in aspect 9, wherein the base further comprises a fluid conduit whereby fluid can enter the bearing.

Aspect 11. The tensioner of aspect 9, wherein the pulley is journaled on a bearing.

Aspect 12. The tensioner of aspect 11, wherein the bearing comprises a needle bearing.

Although the present invention and its advantages have been described in detail, it should be understood that various changes, substitutions and alterations can be made herein without departing from the scope of the invention as defined by the appended claims. Moreover, the scope of the present application is not intended to be limited to the particular embodiments of the process, machine, manufacture, composition of matter, means, methods and steps described in the specification. As one of ordinary skill in the art will readily appreciate from the disclosure of the present invention, processes, machines, manufacture, compositions of matter, means, methods, or steps, presently existing or later to be developed that perform substantially the same function or achieve substantially the same result as the corresponding embodiments described herein may be utilized according to the present invention. Accordingly, the appended claims are intended to include within their scope such processes, machines, manufacture, compositions of matter, means, methods, or steps. The invention disclosed herein may suitably be practiced in the absence of any element that is not specifically disclosed herein.

Claims (12)

1. A synchronous belt drive system comprising:

a synchronous belt (200) having a tensile cord comprising high modulus fibers;

a drive sprocket and at least one driven sprocket, at least one of the drive sprocket and the at least one driven sprocket being an oblong sprocket (10); and

a tensioner, said tensioner comprising: a base having an axially extending cylindrical portion including a radially outer surface and a receiving portion radially inward of the radially outer surface; an eccentric arm pivotally engaged with the radially outer surface; a torsion spring disposed within the radially inner receiving portion, the torsion spring applying a biasing force to the eccentric arm; and a pulley journalled to the eccentric arm.

2. The synchronous belt drive system of claim 1, wherein the eccentric arm, the pulley, and the torsion spring are concentrically arranged such that none of the eccentric arm, the pulley, or the torsion spring is axially displaced from one another along axis a-a.

3. The synchronous belt drive system of claim 1, wherein the obround sprocket (10) comprises a toothed surface and at least one linear portion (16) disposed between two arcuate portions (14, 15) having constant radii (R1, R2), the linear portion having a predetermined length.

4. The synchronous belt drive system as in claim 1, wherein the amplitude and phase of the oblong sprocket (10) is such that the angular displacement timing error between the drive sprocket and the driven sprocket is less than 1.5 degrees peak-to-peak.

5. The synchronous belt drive system of claim 1 wherein the high modulus fiber is one or more selected from the group consisting of glass fiber, PBO, aramid, and carbon fiber.

6. The synchronous belt drive system of claim 1, wherein the high modulus fiber is a high strength glass fiber.

7. The synchronous belt drive system of claim 1, wherein the high modulus fiber is carbon fiber.

8. The synchronous belt drive system of claim 1, wherein the tensile cord is a hybrid cord comprising carbon fiber and glass fiber.

9. The synchronous belt drive system of claim 1, defining an overall system package width of less than 20 mm.

10. The synchronous belt drive system of claim 1, defining an overall system package width of 18mm or less.

11. The synchronous belt drive system of claim 1, defining an overall system package width of 16mm or less.

12. The synchronous belt drive system of claim 1, the synchronous belt having a width of about 14mm or less.

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