JP6995895B2 - Synchronous belt drive system - Google Patents

Description

The present invention generally includes a synchronous belt, an auto tensioner, a drive pulley, and one or more driven pulleys, wherein at least one of these pulleys is oval with respect to a synchronous belt drive system as used in an internal combustion engine. There, and more particularly, relates to a belt drive system that is suitable for traveling in oil and has a package width much narrower than known systems.

Synchronous belt drive systems for most engines are located outside the engine block and at the front of the engine. Synchronous belts may run unlubricated in so-called "dry" belt applications. The result of this proposal is that the drive belt and camshaft must exit the block through the oil seal, which is prone to oil leaks over time. The advantage is that the width of the belt drive mechanism in the engine room can be easily accommodated, that is, there are relatively no restrictions on the width of the drive system. Nevertheless, some engine rooms are limited in space and a reduction in vehicle weight is desired so that the same drive load, tension, and timing requirements can be achieved with the reduced system weight and width. A narrower dry belt drive system is desired. As used herein, "width" refers to dimensions in the axial direction of the drive system or element, i.e., perpendicular to the plane of drive, as viewed in the layout diagram of the drive mechanism.

Internal combustion engine timing chains are generally mounted inside the engine, which facilitates lubrication. Another engine timing proposal is to provide a synchronous belt drive system within the engine block, similar to the design of a timing chain drive mechanism. This proposal reduces the number of seals on the shaft. This proposal requires a lot of constraints on the width of the system to avoid the cost of adapting to the width of the oversized engine block and the timing chain drive mechanism for this engine. In addition, this so-called "wet" belt application requires the belt material to have excellent resistance to oils and other engine fluids. Therefore, a narrower wet belt drive system that directly replaces the timing chain system and copes with the same drive load, tension and timing requirements with reduced system weight is desirable in an environment where the vehicle travels in contact with oil.

As an example, one known commercial in-oil belt synchronous drive system has a belt about 18 mm wide, a circular pulley about 19-23 mm wide, and a conventional tensioner occupying a width of about 35-38 mm. The maximum timing error is typically considered to be about 2 ° at the peak peak value. The mass of the system is said to be about 2500g in total including the VVT element and about 1700g without the VVT.

US Pat. No. 9927001 (Dayco Europe SRL) is representative of this technique. It describes a belt life test using a 19 mm wide belt that runs in contact with oil.

It is desired to reduce the total system width, which is less than 18 mm, i.e. less than half the typical current system width. Doing this requires a somewhat narrow belt and pulley and a fairly narrow tensioner. This is not an easy task. Reducing the belt width increases belt tensile strain, contact pressure, and tooth deformation, which in turn increases timing error and promotes belt deterioration, reducing synchronous system performance and life expectancy. Because. What is desired is to reduce the system width to 20 mm or less and the weight accordingly, while reducing the timing error by about 50% without losing belt life.

The present invention is directed to systems and methods that provide narrower, wet or dry belt synchronization systems that can address common drive loads and tensions and timing requirements with reduced system weight and width. ..

The present invention relates to a system having a highly elastic synchronous belt, an oval sprocket and an autotensioner having a narrow outer shape. The highly elastic synchronous belt preferably has a tension core of the highly elastic fiber, such as containing the highly elastic glass fiber, carbon fiber, PBO or aramid, and a mixture thereof. The oval sprocket has a phase and magnitude selected so that the timing error is reduced to 1.5 ° at the peak peak value, preferably 1.0 ° or less at the peak peak value. The oval sprocket has a toothed surface and at least one straight portion arranged between two arc portions having a fixed radius, the straight portion having a predetermined length. The tensioner is preferably in the range of 100N to 600N and has damping and tensioning features selected to maintain loosening belt tension, preferably with torsion springs and asymmetric damping. The tensioner is preferably disposed within the receptacle with a base having a radial outer peripheral surface and a receptacle and having an axially extending cylindrical portion, and an eccentric arm swingably engaged with the radial outer peripheral surface. A torsion spring is provided, and the torsion spring applies an urging force to the eccentric arm and is provided with a pulley pivotally supported by the eccentric arm. Preferably, none of the eccentric arm, the pulley and the torsion spring is displaced axially along the axes AA from the other.

The above is rather a broad overview of the features and technical advantages of the invention in order to better understand the detailed description of the invention below. Further features and advantages of the invention are described below, which form the subject matter of the claims of the invention. It should be understood by those skilled in the art that the disclosed concepts and special embodiments are readily available on the basis of modifying or designing other configurations to achieve the same object as the present invention. be. It should also be understood by those skilled in the art that such an even configuration does not deviate from the scope of the invention described in the appended claims. The novel features that appear to be features of the present invention, along with further objectives and advantages, regarding the method of construction and method of action, will be better understood from the description below when considered in connection with the accompanying drawings. To. However, it should be clearly understood that each feature is for illustration and description purposes only and is not intended as a definition of the limitations of the invention.

The accompanying drawings, which are incorporated in this specification and constitute a part thereof, in which similar numbers indicate similar parts, are used to show a preferred embodiment of the present invention and to explain the principle of the present invention together with the description.

It is a side view of an oval sprocket.

It is a side view of the oval sprocket of another embodiment.

It is a perspective view of a twin cam, an in-line 4-cylinder, 4-stroke gasoline engine.

It is a perspective view of a single cam, an in-line 4-cylinder, 4-stroke diesel-driven engine in which a fuel pump is incorporated behind a camshaft.

FIG. 3 is a perspective view of a single-cam, 4-cylinder, 4-stroke diesel-driven engine in which a fuel pump is incorporated into a synchronous belt drive system.

It is a schematic diagram of a twin-cam, 4-cylinder, 4-stroke gasoline-driven engine.

It represents the general total load characteristics for a driven sprocket of a 4-cylinder, 4-stroke diesel engine and includes the extracted 1.5th and 2nd order vibration curves.

It represents the secondary vibration load characteristics for the drive sprocket of a 4-cylinder, 4-stroke engine.

It represents the load characteristics of the 1.5th order for the drive sprocket of a 4-cylinder, 4-stroke common rail diesel engine equipped with a 3-piston fuel pump (or other device that induces 1.5th-order vibration).

A group of curves representing the stress / strain relationship in a synchronous belt.

It is a series of curves showing the effect of fading / misfading of the oval sprocket on the vibration of the engine in the system of FIG.

It is a graph which shows the angular vibration characteristic in the camshaft of the engine shown in FIG. 6 before and after applying an oval sprocket.

It is a graph which shows the tension side tension characteristic of an engine shown in FIG. 6 before and after applying an oval sprocket.

It is a graph of the angular vibration with respect to the rotation speed of a crankshaft.

It is a graph of the angular vibration with respect to the rotation speed of the crankshaft in the intake cam.

It is a graph of the angular vibration with respect to the rotation speed of the crankshaft in the exhaust cam.

It is a graph of the angular displacement with respect to the rotation speed of the crankshaft in the intake cam.

It is a graph of the angular displacement with respect to the rotation speed of the crankshaft in the exhaust cam.

It is a graph which shows the effect of fading of an oval sprocket with respect to the timing error of each camshaft.

It is a graph which shows the effect of fading of an oval sprocket on the timing error of each camshaft which has a standard and a highly elastic belt.

It is a graph which shows the effect of an oval sprocket on the timing error by a belt width.

It is a graph which shows the effect of an oval sprocket on the timing error by the magnitude of eccentricity.

It is an exploded view of a tensioner.

It is an exploded view seen from above the tensioner.

It is a perspective view of the base of a tensioner.

It is a perspective view of the eccentric arm of a tensioner.

It is a perspective view of the torsion spring of a tensioner.

It is sectional drawing of a tensioner.

It is an exploded view of another tensioner.

FIG. 29 is a view seen from above the other tensioners.

FIG. 29 is a cross-sectional view of another tensioner in FIG. 29.

It is a side view of another tensioner.

It is a perspective view of another tensioner of FIG. 32.

It is a partial cross-sectional view of a synchronization belt.

FIG. 3 is a diagram of a synchronous belt drive system used to test the features of the present invention.

The present invention relates to a narrowly designed synchronous belt drive system that includes a highly elastic synchronous belt and one or more oval sprockets and an auto tensioner. The highly elastic synchronous belt preferably has a conventional glass fiber tension member, but has a tensile elasticity about twice the elasticity of a synchronous belt of similar width. The highly elastic synchronous belt is preferably made of a heat resistant and oil resistant material. The oval sprocket preferably has a magnitude and phase such that the angular displacement timing error between the sprocket and the oval sprocket is less than 1.5 ° at the peak peak value. The autotensioner has a radial outer peripheral surface and a receiving portion that is radially inside of the radial outer peripheral surface, and swingably engages with a base having a cylindrical portion extending in the axial direction with the radial outer peripheral surface. An eccentric arm, a torsion spring arranged in a receiving portion on the inner side in the radial direction and applying an urging force to the eccentric arm, and a pulley pivotally supported by the eccentric arm may be provided. Preferably, the eccentric arm, the pulley and the torsion spring are coaxially arranged so that none of the eccentric arm, the pulley or the torsion spring is axially displaced along the axis AA from the eccentric arm, the pulley or the torsion spring. It is placed and has the minimum width.

High elasticity synchronous belt

The configuration of the sync belt is shown in FIG. The belt 200 has teeth 214 alternating with land portions 212 on one side and a flat back surface 220. The body rubber or elastomer includes a tooth rubber 212 and a back rubber 222. The tooth side is covered with the tooth cloth 216, and the dorsal side 220 is covered with the back cloth 224. The repeating length of the tooth is pitch P. The tension member 218 is embedded in the rubber of the belt body to impart high elasticity to the belt.

The tooth cloth 216 includes a canvas and one or more treatment agents for enhancing one or more belt performances such as adhesion, oil resistance, wear resistance, etc., as well as system performance such as timing error and durability. The treatment agent for canvas may be a suitable treatment agent known in the past. Similarly, the back cloth 224 contains a canvas and one or more treatment agents that are the same as or different from the tooth cloth. Thus, the phrase "cloth" is used to describe a canvas with one or more treatment agents, including those that are prepared to be incorporated into the belt. "Canvas" usually refers to the material of the woven, non-woven or knitted fabric before it has been treated.

The tooth cloth may be a suitable woven cloth, knitted cloth, or non-woven fabric having appropriate elongation, strength, abrasion resistance, heat resistance, and environmental resistance, as required for the application. For use in the flow-through belt manufacturing process (defined below), it is preferred that the elongation in the length direction is greater than 80% or 100%. Due to the preforming process, smaller elongations may be appropriate, or no significant elongations may be appropriate. The canvas may contain highly elastic, oil resistant, abrasion resistant, heat resistant fibers such as nylon, aramid, PPS, PEEK, polyester, and combinations thereof. Sufficient stretch yarns are obtained by known suitable methods including weaving, wrapping around elastic cores, biasing, and combinations thereof.

The backing fabric may be a suitable woven, knitted or non-woven fabric having appropriate elongation, strength, abrasion resistance, heat resistance (high temperature or low temperature), and environmental resistance for the application. Generally, it is laid flat on the back of the belt, so there is no special requirement for the amount of stretch of the backing cloth. Some elongation is preferred to maintain the flexibility of the belt. Back fabrics have been shown to improve cold tolerance and reduce back cracking during repeated cold starts.

Preferred canvas treatment agents include epoxy or epoxy rubber treatment agents as described in US Patent Application Publication No. 2014/0080647 by Yamada et al., Which publication is incorporated herein by reference. Such treatment agents improve the wear resistance, oil resistance and water resistance of the tooth cloth and have sufficient durability even when used under high temperature and high load conditions or in oil or water environment. Intended to provide a toothed belt.

Suitable rubber compounds can be used for the tooth rubber 212 or the back rubber 222. Further, there may be another rubber layer, if necessary, such as an adhesive rubber layer or another layer that comes into contact with the tension core wire 218. If desired, the same or different composites may be used in the teeth, in the tension core layer, on the dorsal side of the belt and elsewhere. U.S. Pat. No. 6,358,171, granted to Whitfield, is incorporated herein by reference and describes a representative rubber compound for tooth rubber or belt body rubber. As described herein, the rubber body rubber compound may contain a nitrile group containing a copolymer rubber such as HNBR, and the rubber may also contain a third monomer that reduces the glass transition of the rubber. good. The rubber composite may contain from about 0.5 to about 50 parts by weight (PHR) of fiber reinforced material relative to the rubber. U.S. Pat. No. 9,140,329 is incorporated herein by reference and describes other representative rubber compounds for tooth rubber or belt body rubber. As described herein, the belt body rubber compound may include HNBR or HXNBR rubber, resorcinol, and melamine compounds.

The rubber synthetic material of the belt body may further contain known additives such as a filler, a plasticizer, a deterioration inhibitor, a processing aid, a vulcanizing agent, an auxiliary agent, and a phase solvent.

The tension core wire 218 of the belt may be known, but is preferably fiberglass, PBO, aramid, carbon fiber, or a combination thereof of two or more. The tension core preferably contains an adhesive treatment agent that is highly resistant to oil for use in an oil contact environment. For example, the adhesive treatment agent may be based on a nitrile-containing latex or rubber, or other oil resistant material. Preferred tension cores are carbon fibers as disclosed in US Pat. No. 6,945,891 to Natson, or glass / carbon as described in US Pat. No. 7,682,274 to Akiyama et al. It is a mixed core wire. Preferred fiberglass for tension cores is high-strength fiberglass such as K glass, U glass, M glass, or S glass.

The toothed belt of the present invention may be manufactured by a known belt manufacturing method. The most common proposal is to wrap various materials around the grooved mandrel, first with the tooth cover jacket, then with the tension core and body rubber, and finally with the back jacket. Then, along with the belt slab, the mandrel is carried into a pressure shell that is heated and pressed to narrow it down together with the material, and the rubber flows into the tooth groove and pushes the tooth cover jacket into the shape of the groove (“flow-through type”). Known as). Alternatively, the tooth cover jacket may be preformed into an approximate groove shape, with rubber selectively filling the teeth before or during the placement of the tooth cover jacket on the mandrel (preforming method). Other variants of these methods are possible. A major additional feature of manufacturing a back canvas belt is to obtain the desired final belt thickness, as the back of the belt cannot be cut for dimensioning when a belt with rubber on the back is created. The rubber layer must be measured with great care.

Rubber compositions generally have a level of elasticity that contributes sufficiently to the tooth stiffness and load bearing of the tooth and synchronous belt. Similarly, the tooth cover jacket helps to reinforce the teeth, which also helps to stiffen the teeth and withstand the load of the belt. Tension cores generally dominate the tension performance of synchronous belts, such as elasticity (or span stiffness) and strength. Optimizing the combination of belt-spun stiffness and tooth stiffness by selecting these materials reduces the load on the system and minimizes the strength required for the belt, while narrower belts have the same system timing error. Was found to have. Increasing the span stiffness is particularly desirable for conventional belts with glass fiber cores. According to computer simulations, a series of belts with up to twice the span stiffness and the same tooth stiffness have the same system timing error, but with the system's maximum belt tension, maximum belt effective tension and maximum tensioner span tension. To reduce.

Oval sprocket

The present invention is a synchronous belt drive system, the synchronous belt drive system having a first length having a toothed surface and at least one straight portion (16) disposed between two arc portions (14, 15). The circular sprocket (10) is provided, the arc portion has a fixed radius (R1, R2), the straight portion has a predetermined length, and the sprocket (300) has a toothed surface, and the sprocket is an endless toothed member. Engaged with the first oval sprocket by (200), the first oval sprocket (10) has an angular displacement timing error between the sprocket and the first oval sprocket of more than 1.5 degrees at the peak peak value. It has a small size and phase.

FIG. 1 is a side view of an oval sprocket. The sprocket 10 of the present invention includes a toothed surface 11. The toothed surface 11 engages with the toothed belt. The toothed surface 11 includes a land portion 12 and an adjacent groove portion 13. The groove portion 13 has a shape suitable for the tooth shape of the corresponding toothed belt. Toothed belts are also called synchronous belts because they are used to synchronize the rotation of the driven and driven sprockets.

The sprocket 10 comprises a portion 14 and a portion 15. The portion 14 has an arcuate toothed surface 11a with a fixed radius R2. The portion 15 has an arcuate toothed surface 11b with a fixed radius R1. Since the radii R1 and R2 are equally constant, the portions 14 and 15 are arcs of one circle. The use of arcs in this way reduces the complexity of the design and manufacturing process of the sprocket of the present invention.

A straight line portion 16 is arranged between the portions 14 and 15. The portion 16 comprises a rectangular portion having the effect of displacing the portions 14, 15 with respect to each other to form an oval sprocket. The sprocket surface 11 is straight, i.e. straight or flat, between points 160 and 161 and between points 162 and 163.

The flat portion 16 has a length related to the amplitude of torque fluctuations in the system. In this embodiment, the portion 16 has a dimension (W) of about 2 mm between points 160 and 161 and between points 162 and 163. Therefore, the center of curvature 17 of the portion 14 deviates from the center of rotation 20 of the sprocket by a distance W / 2, about 1 mm, which is called the "magnitude" of the eccentricity of the oval sprocket. Further, the center of curvature 18 of the portion 15 is deviated from the center of rotation 20 of the sprocket by a distance of W / 2, about 1 mm. The dimensions shown here are for illustration purposes only and are not intended to be limiting. Therefore, the major axis length of the sprocket has the following dimensions.
Lmajor = R1 + R2 + W
The main length (MG) of each portion 14, 15 has the following dimensions.
MG = (R1 + W / 2) or (R2 + W / 2)
The minor axis length of the sprocket has the following dimensions.
Lminor = R1 + R2
A more common size that may be useful for sprockets with two or more lobes is the difference between the maximum major length and the minimum minor axis length (ie MG-R1 or MG-R2 for these two lobes). Is. For symmetrical two lobes, the size is just W / 2. If you have an asymmetric design or lobe larger than 2, there is a small displacement from W / 2.

The length (W) of the portion 16 is determined by the radii of the portions 14, 15 and depends on the dynamic angular vibration characteristics, which are offset as described elsewhere in this specification. The sprocket 10 can be designed using a fixed surface pitch, a fixed angle pitch, or a combination of the two. "Surface pitch" is defined as the distance between any two consecutive corresponding "pitch points" on the sprocket's OD, measured along the OD line.
The fixed surface pitch is calculated as follows.
SP = (((((Ng × Nom Pitch) / Pi) -PLD) × Pi) / Ng)
Here, SP = surface pitch Ng = number of grooves in the sprocket Nom Pitch = nominal system pitch Pi = about 3.141
PLD = system diameter PLD
An "angular pitch" is defined by any successive angular difference corresponding to a "pitch point" on the sprocket and is measured in degrees or radians. The fixed angle pitch is defined as follows.
AP = 360 / Ng (degrees)
here,
AP = Angle pitch Ng = Number of sprocket grooves

The sprocket groove contours may be individually designed to suit the particular behavior of the engine.

In the combination of tooth modulus and sprocket offset (W / 2), the beltspun modulus is optimized to offset tension fluctuations at a particular engine speed. Therefore, in this application, the belt is designed to carry the required tension load and is analyzed and designed as a spring member of the system. The dynamic response of the system is an interaction that achieves a combination of belt modulus and oval sprocket radius (R1 and R2) that offsets all tension fluctuations so that they are not substantially transmitted through the belt and belt drive system. Selected by the process.

FIG. 2 is a side view of another embodiment of the sprocket. This embodiment is different from that shown in FIG. 1 and includes three straight lines arranged between the arcuate portions 14, 15 and 16. The three straight portions (161 to 162), (163 to 164), and (165 to 166) are arranged between the arc portions 14, 15, and 16. Each of the arc portions 14, 15 and 16 has constant and equal radii R1, R2 and R3, respectively. The three straight portions are evenly distributed on the peripheral edge of the sprocket at intervals of about 120 °. FIG. 9 is a representative example of the load characteristic of the 1.5th order vibration in the system using the sprocket shown in FIG.

FIGS. 3, 4, and 5 are layouts of some typical drives of a 4-cylinder, 4-stroke internal combustion engine using a toothed belt system to drive a camshaft and accessories. These engines generally exhibit high secondary vibrations. Depending on the specifications of the fuel pump, some diesel engines contain 1.5th order vibration, which is dominant. Graphs showing such vibrations are shown in FIGS. 7, 8 and 9.

To offset the secondary vibration, the sprocket 10 of the present invention is attached to the crankshaft CRK of the engine. Depending on the existence of other dominant orders, it may be necessary to apply sprockets of other embodiments. These may be mounted on the crankshaft, but may be similarly applied elsewhere in the system, such as water pumps, fuel pumps, or camshaft sprockets. The engine shaft is the drive unit of the entire belt drive system. The driving direction of the belt is DoR. Due to the sprocket ratio, the engine crankshaft CRK makes two revolutions per revolution of the camshaft CAM1.

In FIG. 3, the sprocket 300 is connected to the camshaft CAM1 and the sprocket 304 is connected to the second camshaft CAM2. Conventionally known idlers IDR1 and IDR2 are used to maintain proper belt path and tension control. The sprocket 100 is connected to the water pump WP. The belt 200 is hung between several sprockets. The direction of rotation of the belt 200 is shown as DoR. The position where the belt 200 engages with the crankshaft sprocket CRK is 201. The inertia and torque load of the camshaft is represented by 301.

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

Similarly, in FIGS. 4 and 5, the camshaft sprocket 300 is attached to the engine camshaft CAM. In FIG. 4, the load characteristic 301 includes the torque characteristic of the fuel pump mounted on the rear side of the camshaft, whereas in FIG. 5, the torque of the fuel pump is represented by the load characteristic 302. Inertia and torque loads (301, 302, 101) caused by other components such as water pumps and vacuum pumps may be present as well. That is, it is WP (101) in FIGS. 4 and 5. In FIG. 4, IDR1 and IDR2 are known idlers for properly guiding the belt 200. In FIG. 4, the belt span length between the crankshaft sprocket 10 and the cam sprocket 300 is SL.

In gasoline engines, the dominant periodic variable torque load is usually a characteristic of camshafts. In a diesel engine, the dominant order may be caused by the camshaft and / or fuel injection pump that will be included in the drive system. The torque generated by the water and vacuum pumps varies, but they are not periodic, they are essentially the same period or frequency as the camshaft, and are usually not the dominant characteristic of drive vibration.

FIG. 5 is a perspective view of another single cam engine embodiment in which the drive device of the diesel engine includes a fuel injection pump. In this embodiment, in addition to the system shown in FIG. 4, the system further comprises a sprocket 305 coupled to the fuel pump IP. Also shown is a sprocket P1 that is engageable with other multi-rib belts used to drive various engine accessories (not shown). In FIG. 5, the cam load is indicated by 301 and the fuel pump load is indicated by 302. The sprocket 100 is connected to the water pump WP. In FIG. 5, the torque load generated by the fuel injection pump is indicated by 302.

The general full load characteristics of a 4-cylinder, 4-stroke engine are shown by curve E in FIG. Curves D and C are general 2nd and 1.5th order vibration characteristics, extracted from the total load characteristics. The load characteristics of an in-line 4-cylinder, 4-stroke, gasoline-powered engine usually do not include vibrations of degree 1.5.

The change in the average radius of the sprocket 10 at the belt engagement point 201 when the sprocket 10 of the present invention rotates is shown by the curve C in FIGS. 8 and 9. The integral of the curve C, which is the change in the effective length of the belt in FIG. 4, is the curve D in FIGS. 8 and 9. The derivative of the change in average sprocket radius is the acceleration of the point given on the toothed surface 11 by the change in sprocket shape.

In order to cancel the secondary vibration, the flat portion 16 of the elliptical sprocket 10 is arranged in relation to the camshaft sprocket 300 in terms of timing. That is, for example, the effective length of the belt 200 between the sprocket 300 and the sprocket 10 in FIG. 4 changes so as to substantially cancel the alternately fluctuating belt tension caused by the periodic torque fluctuation of the camshaft. Set. As an example of a design that cancels secondary vibration, it is achieved by matching the maximum length of the sprocket 10 (R1 + R2 + W) with the entry point 201 of the belt when the camshaft torque, that is, the belt tension, is maximized. can do.

The complete dimensional characteristics of the drive, including the oval sprocket, are parameters such as variable torque, beltspun modulus, inertia of each driven accessory in the system, belt mounting tension, and interaction between the belt and sprocket. Depends on. The interaction between the belt and the sprocket depends on parameters such as the number of teeth meshing in the sprocket, the elastic modulus of the belt teeth, the belt dimensions, and the coefficient of friction between the belt and the sprocket surface.

FIG. 6 is a schematic diagram of a twin cam, 4-cylinder, 4-stroke gasoline engine. The illustrated system comprises cams CM1, CM2 and a belt B hung between them. It also comprises a tensioner TEN, a water pump WP, and a crankshaft sprocket CRK. The rotation direction of the belt B is indicated by DoR. Important span lengths are between the sprocket CRK and the sprocket IDR, the sprocket IDR and the sprocket WP, and the sprocket CRK and the sprocket WP. In FIG. 6, the belt span length between the crankshaft sprocket CRK and the cam sprocket CM1 is SL. Since there is no large load impact between CM1 and CRK in the DoR direction, these can be treated as one span SL in the calculation. The general approximations of the system variables shown in FIG. 6 are as follows.
General cam torque fluctuation: + 40N / -30N
Belt span elastic modulus: 240Mpa

The value of inertia of a general component is CRK = 0.4 gm ²
CM1 = CM2 = 1.02gm ²
WP = 0.15gm ²

Belt mounting tension: 400N (mounting tension is maintained by the tensioner TEN by a known method)

Teeth engaged with 3 sprockets: CRK ⇒ 9 teeth; CM1, CM2 ⇒ 15 teeth

Belt dimensions: width = 25.4 mm; length = 1257.3 mm

The general value of the coefficient of friction of the sprocket surface 11 is in the range of 0.15 to 0.5, typically 0.2.

Typical belt mounting tensions can range from 75N to 900N according to system requirements.

The belt span elastic modulus depends on the structure of the tension member, the number of strands of the tension member in the belt, and the belt width. An example of a belt span elastic modulus for a belt width of 25.4 mm with 20 tension members is around about 240 Mpa.

FIG. 7 represents typical full load characteristics for a driven sprocket of a 4-cylinder, 4-stroke diesel engine and includes the extracted 1.5th order vibration (curve C) and second order vibration (curve D) curves. The load characteristics of an in-line 4-cylinder, 4-stroke gasoline-powered engine usually do not include 1.5th-order vibration. The offset corresponds to W / 2. The full load corresponds to curve E in FIG.

In FIG. 7, line A has 0 torque. Line B shows the average torque in the belt drive system. The curve C shows the torque characteristics of the 1.5th-order vibration extracted from the full load curve E. The curve D is the torque characteristic of the secondary vibration extracted from the total load curve E. Curve E is the total torque characteristic of the engine measured at the crankshaft CRK. The area below the curve E shows the work used to rotate the engine at a particular speed.

FIG. 8 shows the secondary vibration load characteristics (curve B) of the drive sprocket of the 4-cylinder, 4-stroke engine including the oval sprocket and the change in radius (curve C) due to the change in belt span length (curve D). show.

In FIG. 8, the straight line A has a torque of 0. The curve B is the torque characteristic of the secondary vibration extracted from the total load. Curve C is a change in the effective radius of the crankshaft pulley caused by segment 16 in FIG. 1 when the crankshaft pulley rotates 360 °. Curve D is an integral of curve C, which is an effective change in belt drive span length caused by the sprocket shown in FIG.

FIG. 9 shows the torque characteristic B of the 1.5th order vibration with respect to the drive sprocket of a 4-cylinder, 4-stroke diesel engine equipped with a 3-piston type fuel pump (or another device that induces the 1.5th order vibration). Includes a change in sprocket radius length (curve C) of an elliptical sprocket (FIG. 2), an embodiment having three alternative lobes, and a consequent change in belt span length (curve D). There is. The belt span length is, for example, the distance between the cam sprocket CAM and the crankshaft sprocket CRK in FIG.

In FIG. 9, the straight line A has a torque of 0. Curve B is the torque characteristic of the 1.5th order vibration extracted from the total load. The curve C is a change in the effective radius of the crankshaft pulley when the crankshaft pulley rotates 360 °. Curve D is an integral of curve C and is an effective change in drive length caused by the alternative embodiment shown in FIG.

The elastic moduli of the tension members of various belts used in the system of the present invention are shown in FIG. Curves SS1 to SS6 are known as stress-strain curves for various belts 200. Each curve shows the elastic modulus when different materials are used for the tension core wire in the belt. The elastomeric HNBR belt body is exemplary and not limiting. In addition to HNBR, other belt body materials may include EPDM, CR (chloroprene), polyurethane, or a combination of two or more of those described above. The material includes:
SS1 (fiberglass # 1 tension core wire, HNBR body)
SS2 (fiberglass # 2 tension core wire, HNBR body)
SS3 (fiberglass # 3 tension core wire, HNBR body)
SS4 (carbon fiber tension core wire, HNBR body)
SS5 (Aramid tension core wire, HNBR body)
SS6 (carbon fiber tensile strength core wire, HNBR body)

As is known, the elastic modulus of each tension member is the slope of each curve SS1 to SS6. In general, this measurement and calculation is done on a substantially straight line portion of the curve. In addition to fiberglass, carbon fiber, and aramid, other tension members may also include fine stainless steel filament wires or PBOs.

M = Δstress / Δstrain (measured in the substantially straight line of the curve)

The elastic modulus of the belt span depends on the structure of the tension member, i.e., the number of strands of the tension member in the belt and the belt width. An example of the belt span elastic modulus with respect to the curve SS1 is about 242 MPa for a belt having a width of 25.4 mm and having 20 strands of the fiberglass tension member.

FIG. 11 is a set of curves showing the effect of fading / misfading of the major axis length of the elliptical sprocket on engine vibration in the system shown in FIG. Curve D is the optimal timing arrangement between the position of the sprocket major axis length and the torque pulse with respect to the belt entry point 201. The curves A, B, and C are those whose timings are shifted clockwise by +6, +4, and +2 teeth from the position of the curve D, respectively. The curve E is obtained by shifting the timing by two teeth counterclockwise. The fading of the maximum belt span length with respect to the maximum torque and inertial load varies depending on the dominant vibration order in the drive, which should be reduced by the system. Belt entry point 201 is the point at which the belt engages the sprocket. In FIG. 3, the span length is SL.

With respect to angular spacing or fading, the allowable angular tolerance is calculated using the following equation.
± (360/2 x number of sprocket grooves)

The belt drive span length is maximized when the torque is maximum.

FIG. 12 is a graph showing the effect of an elliptical sprocket with exactly fading provided on the twin cam, 4-cylinder, 4-stroke engine shown in FIG. Curves A and B show measured values of angular vibration in each of the intake and exhaust camshaft sprocket in a prior art design using a circular sprocket.

For comparison, curves C and D show measured values of angular vibration in each of the intake and exhaust camshaft sprockets when the sprocket of the present invention was used for the crankshaft. The total damping in angular vibration is about 50%.

Similarly, FIG. 13 is a graph showing the effect of an precisely faded oval sprocket as shown in FIG. 1 on the twin cam, 4-cylinder, 4-stroke engine shown in FIG. Curves A, B, and C each show maximum, average, and minimum tension side tension measurements over a range of engine speeds in the prior art. In this example, the tension was measured at the IDR position in FIG. To extend the life of the belt, the tension on the belt tension side must be minimized. Curves D, E, and F show the maximum, average, and minimum belt tension side tension measurements when the sprocket of the present invention was used. In the resonance speed range of the engine (about 4000 rpm to about 4800 rpm), the total damping at the tension side mounting tension is in the range of 50-60%. Damping at belt tension side tension indicates a significant improvement in the lifespan in which the belt is used.

The system of the present invention is useful for reducing timing errors in IC engines. Timing errors are positional mismatches between driven and driven axes caused by random elements such as vibrations, element inaccuracies, and elastic strains. In this case, it is the rotational inaccuracy of the camshaft (driven) of the IC engine compared to the crankshaft (drive) of the engine. It is usually reported as the peak peak value of the angle. For example, referring to FIG. 3, the sprocket 300 and the sprocket 304 are oval, respectively. The use of oval sprockets significantly reduces timing errors, which in turn results in improvements in fuel economy, low emissions, and generally improves engine performance and efficiency. At the element level, reducing timing errors and reducing system load leads to better durability and lower NVH problems. The reduction in tension reduces NVH levels, especially meshing vibrations, in the drive. The application of oval sprockets to reduce timing errors is not limited to engine camshafts. The benefits are similarly obtained by placing the oval sprocket on the crank or fuel pump.

FIG. 14 is a graph of angular vibration with respect to the rotational speed of the crankshaft. The typical angular frequency decreases as the engine speed increases. FIG. 14 shows data for an electric engine and a combustion engine. In an electric engine, the crankshaft is driven by an electric motor and there is no fuel combustion in each cylinder. In a combustion engine, the crankshaft is driven by burning fuel in each cylinder in the usual way in an internal combustion engine. The electric engine (MER) produces less angular vibration than the combustion engine (FER) for a given engine speed.

FIG. 15 is a graph of the angular vibration of the intake cam with respect to the rotation speed of the crankshaft. The oval sprocket is attached to the intake valve train camshaft. Three states are shown. The first is a standard drive system without oval sprockets (curve A). The second has an elliptical sprocket (curve B), and the third has an elliptical sprocket and a highly elastic belt (curve C). The phase and size of the oval sprocket is 1.5 mm at 10.5 pitch from the 3 o'clock position. The elastic modulus of the standard belt is 630,000 N, and the elastic modulus of the high elastic belt is 902,000 N. The modulus of elasticity is expressed in Newton (N) and is defined as the force required to extend the unit length by 100%.

The angular vibration (curve C) in the third state is significantly smaller than 0.5 degrees at the peak peak value when compared to the value for the standard drive system at about 1.5 degrees. It was reduced and both were measured at 4000 RPM.

FIG. 16 is a graph of the angular vibration of the exhaust cam with respect to the rotation speed of the crankshaft. The oval sprocket is attached to the exhaust valve train camshaft. Three states are shown. The first is a standard drive system without oval sprockets (curve A). The second has an elliptical sprocket, and the third has an elliptical sprocket and a highly elastic belt (curve B). The vibration angle in the third state is significantly reduced to about 0.5 degrees at the peak peak value when compared to the value for the standard drive system at the peak peak value of about 1.5 degrees, both measured at 4000 RPM. (Curve C). However, depending on the engine, the improvement ranges from 1.5 degrees to 0.5 degrees with peak peak values, and the reduction is only over 60%. The phase and size of the oval sprocket is 1.5 mm at 23.5 pitches from the 3 o'clock position. The elastic modulus of the standard belt is about 630,000 N, and the elastic modulus of the high elastic belt is 902,000 N.

FIG. 17 is a graph of the angular displacement of the intake cam with respect to the rotational speed of the crankshaft. Angle displacement, also called timing error, is measured with respect to the position of the crankshaft. The oval sprocket is attached to the intake valve train camshaft. Three states are shown. The first is a standard drive system without oval sprockets (curve A). The second has an elliptical sprocket, and the third has an elliptical sprocket (curve B) and a highly elastic belt (curve C). The angular displacement in the third state is significantly reduced until the peak peak value is less than about 0.5 degrees when compared to the value for the standard drive system at about 1.5 degrees, both. Measured at 4000 RPM (curve C). However, depending on the engine, the improvement ranges from below 1.5 degrees to about 0.5 degrees at peak peak values, and the reduction is only over 60%. The phase and size of the oval sprocket is 1.5 mm at 10.5 pitch from the 3 o'clock position. The elastic modulus of the standard belt is about 630,000 N, and the elastic modulus of the high elastic belt is about 902,000 N.

FIG. 18 is a graph of the angular displacement of the exhaust cam with respect to the rotational speed of the crankshaft. The oval sprocket is attached to the exhaust valve train camshaft. Three states are shown. The first is a standard drive system without oval sprockets (curve A). The second has an elliptical sprocket (curve B), and the third has an elliptical sprocket and a highly elastic belt (curve C). The angular displacement in the third state is significantly reduced to about 0.5 degrees at the peak peak value when compared to the value for the standard drive system at about 1.5 degrees at the peak peak value, both measured at 4000 RPM. Was done. However, depending on the engine, the improvement ranges from below 1.5 degrees to about 0.5 degrees at peak peak values, and the reduction is only over 60%. The phase and size of the oval sprocket is 1.5 mm at 23.5 pitches from the 3 o'clock position. The elastic modulus of the standard belt is about 630,000 N, and the elastic modulus of the high elastic belt is about 902,000 N.

FIG. 19 is a graph showing the effect of fading of the oval sprocket on the timing error of each camshaft. The Y-axis is the angular displacement or timing error of each camshaft with respect to the crankshaft. It is shown as the peak-peak value, that is, the numerical difference between the minimum and the maximum. Columns 1 and 2 of the graph show standard drives configured using circular sprockets. Column 3 shows the use of a tertiary oscillating oval sprocket mounted on the intake and exhaust camshafts. Each sprocket is positioned so that the maximum displacement matches the protrusion of the camshaft. Rows 4 to 13 show the results of various experiments using different oval sprockets. The 3 o'clock position is a reference for all deviations. The numbers given are simply the number of pitches or grooves g in which the sprocket reference point has rotated from that position. The reference point is the point used as a reference for angle measurement. This is set at the 12 o'clock position. CW points clockwise. For example, "EX 23.5g cw" refers to the 3 o'clock position and the exhaust cam oval sprocket having a 23.5 groove shift clockwise from 3 o'clock in the engine.

FIG. 20 is a graph showing the effect of fading of an oval sprocket on the timing error of each camshaft with standard and highly elastic belts. The Y-axis is the angular displacement or timing error of each cam sprocket with respect to the crankshaft in degrees of peak peak value. It is shown as the peak-peak value, that is, the numerical difference between the minimum and the maximum. Columns 1 and 3 of the graph show standard drives configured using circular sprockets. Each row shows the use of a tertiary oscillating oval sprocket mounted on the intake and exhaust camshafts. Each sprocket is positioned so that the maximum displacement matches the protrusion of the camshaft. Rows 2, 4-8 show the results of various experiments using different oval sprockets. The 3 o'clock position is a reference for all deviations. The numbers given are simply the number of pitches or grooves in which the sprocket reference point has rotated from that position. The reference point is the point used as a reference for angle measurement. This is set at the 3 o'clock position. The phase and size of the oval sprocket are 23.5 pitch for exhaust and 10.5 pitch for intake from the 3 o'clock position, 1.5 mm each. The elastic modulus of the standard belt is about 630,000 N, and the elastic modulus of the high elastic belt is about 902,000 N.

FIG. 21 is a graph showing the effect of the oval sprocket on the timing error due to the belt width. Row 1 shows a 14 mm wide belt in a system with circular sprockets. Row 2 shows a 14 mm wide belt in a system with oval sprockets. Row 3 shows a 14 mm wide belt with a highly elastic belt in a system with standard sprockets. Row 4 shows a 14 mm wide belt with a highly elastic belt in a system with oval sprockets. Row 5 shows an 18 mm wide belt with a standard elastic belt in a system with standard sprockets.

FIG. 22 is a graph showing the effect of the oval sprocket on the timing error due to the magnitude of the eccentricity. Each row shows the oval sprockets used for the intake and exhaust camshafts. The magnitude of the eccentricity of each system ranges from 1.0 mm to 1.5 mm.

Tests to evaluate the effectiveness of oval sprockets for reducing vibrations in belt drive systems can be performed on both electric and combustion engines. The result of the improvement of timing error included in the drawing was obtained in the electric engine. In most cases these results transfer to the combustion engine, but in some cases the oval sprocket does not reduce vibration for some engines. Testing should be performed on the combustion engine to ensure that the required improvements have been achieved and are reliable. The steps required to perform the test are known in the field of engine vibration. These also include the need for the vibration sensor to operate in an oil environment, which can withstand up to 160 ° C and can withstand chemical attacks by oils and additives. Consistency checks are performed at the beginning and end of a series of test runs. The measurements are made during the 60 second idle run to the run-up to maximum engine speed. Standard Low-Tech systems are used for data acquisition and analysis.

Tensioner

The present invention comprises a base having a cylindrical portion extending in the axial direction, the cylindrical portion having a radial outer peripheral surface and a receiving portion radially inside the radial outer peripheral surface, and swinging in the radial outer peripheral surface. A tensioner equipped with an eccentric arm that engages freely and a torsion spring that is arranged in a radially inner receiving portion, the torsion spring applies urging force to the eccentric arm, and a pulley that is pivotally supported by the eccentric arm. be.

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

The eccentric arm 420 swings around the cylindrical portion 412. The bush 460 is arranged between the inner peripheral surface 422 and the outer peripheral surface 414. The bush 460 comprises a slot 461 that is substantially aligned with the opening 411 of the cylindrical portion 412. The pulley 440 is pivotally supported on the surface 421 of the needle bearing 450. Needle bearings are used in oily environments. Other conventionally known bearings are also applicable.

The torsion spring 430 engages with the eccentric arm 420 and urges the belt (not shown) to apply the belt load. The end 431 projects from the slot 461 and the opening 411 and engages the receiving portion 424 of the eccentric arm 420. The end 432 engages the receiving portion 415 of the base 410. The torsion spring 430 is generally arranged in the receiving portion 418. The receiving portion 418 is a central recess of the cylindrical portion 412. The torsion spring 430 is coplanar with the bearing 450, the pulley 440 and the eccentric arm 420. The torsion spring 430 is disposed radially inside the pulley 440, bearing 450, bush 460 and cylindrical portion 412. That is, the torsion spring 430, the bearing 450, the pulley 440, and the eccentric arm 420 are all arranged coaxially so that none of these elements is axially displaced along the axes AA from the others. ..

The retaining ring 406 engages the circumferential slot 416 of the base 410. The retaining ring 405 engages with the circumferential slot 423 of the eccentric arm 420. The retaining ring 405 holds the bearing 450 on the eccentric arm 420. The retaining ring 6 holds the eccentric arm 420 to the base 410. The presence of the oil retaining rings 405 and 406 acts as a thrust washer to transmit the axial force.

The pulley 440 is press-fitted into the bearing 450. Fixture 404 projects from the torsion spring 430 and the hole 417 of the base 410 to secure the tensioner 400 to a mounting surface such as an engine (not shown).

The bush 460 has a coefficient of dynamic friction (COF) in the range of about 0.05 to about 0.20. The static COF is preferably smaller than the dynamic COF.

FIG. 24 is an exploded view seen from above. The eccentric arm 420 swings around an axis AA, which axis is coaxial with the cylindrical portion 412 and extends through the fixture 404. The eccentric arm 420 swings around the axes AA. The pulley 440 rotates around B, which is the geometric center of the eccentric arm 420. B is eccentrically offset from axes AA, which allows eccentric swinging motion of the eccentric arm 420, which causes the tensioner 400 to apply a variable load to the belt (not shown). do.

FIG. 25 is a perspective view of the base. The end receiving portion 415 is arranged at one end of the receiving portion 418 of the base 410. The end 432 engages the receiving 415, thereby fixing the end 432 and acting as a reaction point for the torsion spring.

FIG. 26 is a perspective view of the eccentric arm. B is the geometric center of the pulley 420, around which the pulley 420 rotates. The eccentric arm 420 swings around A on axes AA. The receiving portion 424 engages the end portion 431 of the spring 430.

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

FIG. 28 is a cross-sectional view of the tensioner. The torsion spring 430, bush 460, cylindrical portion 412, eccentric arm 420, bearing 450 and pulley 440 are coaxial so that none of these elements are axially displaced along axis AA from the other. Is arranged. This fully nested array minimizes the height (or width) of the tensioner, allowing it to be used in very narrow applications.

FIG. 29 is an exploded view of an alternative embodiment. The elements are the same as those described herein, except that the bearing 451 is a plain bearing and the bush 460 is omitted. This alternative embodiment is configured to run in oil and / or is used in oil splash lubrication. The eccentric arm 420 swings around the axes AA. The pulley 440 swings around the shaft BB. See FIG. 26. The axes AA are located away from the axes BB, thus allowing eccentric swinging motion of the eccentric arm 420 rather than coaxial with the axes AA.

FIG. 30 is a top view of an alternative embodiment of FIG. 29.

FIG. 31 is a cross-sectional view of an alternative embodiment of FIG. The torsion spring 430, the eccentric arm 420, and the bearing 451 are arranged coaxially so that none of these elements is axially displaced along the axis AA from the other. The fluid conduit 471 of the base 410 provides a path through which a fluid such as oil flows from the engine oil system (not shown) to the bearing 451 through the fluid conduit 473 to lubricate the bearing. The O-ring 472 provides a means of sealing the connection to the engine oil system.

FIG. 32 is a side view of an alternative embodiment. Instead of the eccentric arm 420 and the pulley 440, this alternative embodiment comprises a cam 445. The cam 445 operates on the same principle as the eccentric arm 420 and occupies the same position in this device. There is no pulley 445. The cam 445 engages the elongated member 480. The elongated member 480 may be any suitable conventionally known low friction material. The elongated member 480 is also referred to as a slip guide. The chain C slidably engages with the surface of the sliding guide 480. The pivot 481 is arranged at one end of the sliding guide. The sliding guide 480 swings around the pivot 481 in response to the rotation of the cam 445. Due to the eccentric shape of the surface 446, the rotation of the cam 445 swings the sliding guide 480 around the 481 thereby maintaining the load on the chain C. This embodiment is useful, for example, in an internal combustion engine synchronization system. This embodiment can also be used for synchronous belts instead of chains.

33 is a perspective view of an alternative embodiment of FIG. 32. The surface 446 of the cam 445 engages the sliding guide 480.

Synchronous belt drive system

The drive system of the present invention includes a belt, at least one oval sprocket described above, and a tensioner described above. The drive system includes at least two sprockets, a drive sprocket and a driven sprocket. At least one of the driven and driven sprockets is an oval as described herein. The tensioner is preferably designed as described herein and may use a rear idler pulley or slider to contact the belt span.

In one embodiment, the synchronous belt drive system is an internal combustion engine overhead cam drive mechanism, for example, for automobiles or other land vehicles. Examples are introduced above and are shown in FIGS. 3-6. 3 and 6 show a typical dual overhead cam drive system, and FIGS. 4 and 5 show a single overhead cam drive system. These examples have a crank sprocket as a drive unit and one or more cam sprockets as a driven sprocket. Each drive mechanism may also include various rear idler mechanisms and water pumps and the tensioners described herein. The drive mechanism of FIG. 5 includes an injection pump. The drive mechanism may be for a diesel or gasoline engine. In other embodiments, the drive mechanism may be a single driven element such as a pump or a balance shaft including a crank sprocket, a driven sprocket for the driven element, and a tensioner. In each case and innumerable other drive systems, the principles of the invention described herein can travel in the air or in contact with oil or other engine fluids, compared to conventional drive mechanisms. It can be used to provide a synchronous drive system with a substantially narrow package.

The main advantage of the tensioner structure described above for synchronous belt drive systems is compactness. In traditional belt drive mechanisms, the tensioner is often a wide element, significantly limiting the width of the entire package. According to the tensioner structure described here, the belt can be as small as desired, the tensioner and other sprockets can be substantially the same width and have smaller clearances, exist or potential mistakes. It can be slightly wider if desired to adjust the alignment.

By using the oval sprocket together with the highly elastic synchronous belt, it is possible to obtain a drive mechanism that works very well while having a belt with a narrow width. Reducing the belt width is generally associated with increased tension and load (per unit width) in the belt in this art, resulting in span length and tooth deformation, which in turn deteriorates timing accuracy and durability. .. However, the system can show significantly improved timing, along with reduced tension and load on the belt and sufficient durability.

example

In each of the examples below, the test equipment layout is a dual overhead cam drive mechanism for a 3-cylinder engine driven by an electric motor connected to a crankshaft as shown in FIG. The 3-cylinder engine is known as the 1-liter FOX engine manufactured by Ford Motor Company. The standard oil-wet synchronous drive system is modified to accommodate various test drive mechanisms described below. All run in contact with the oil of Castrol Magnet 5W-20 at an oil temperature of 140 ° C. The test equipment layout 240 has a 19-groove drive section or crank pulley 243 and two 38-groove driven cam pulleys 242, 245 with an RPX groove shape and a 9.525 mm pitch to fit the belt 200 with the RPP tooth profile. And include. The speed profile and details of the belt width, pulley and tensioner 244 were changed as described below.

The three representative belts shown in Table 1 were used to test various systems. Belt 1 is a standard belt for a 1 liter FOX engine supplied by Dayco. Belt 2 is an oil belt supplied by Gates Unitta Asia. The belt 3 is an improved product of the belt 2 in which the tension core wire is replaced with a hybrid core wire of carbon fiber coated with U glass fiber (high-strength glass).

Belt 1 appears to be configured as described in WO2005080820 by Deco.

Belt 2 is made of the belt material described in US Patent Application Publication No. 2014/0080647 by Yamada et al., The contents of which are incorporated herein by reference. The tooth cloth is a 2x2 twill weave using para-aramid / nylon elastic stretch yarn for the warp and nylon for the warp. The tooth cloth was treated with a treatment agent of epoxy, NBR latex and a curing agent as described in U.S. Patent Application Publication No. 2014/0080647 of Yamada et al. The treatment was performed by soaking and drying in a conventional oven. The back canvas was a 2x2 twill weave, nylon, stretch fiber treated with NBR latex-based RFL. The rubber material for the belt body (both on the back of the tooth) was based on a nitrile group containing copolymer rubber, which is an HNBR containing short fiber reinforcements, resorcinol, and melamine compounds. The rubber composite of the belt body further contained conventionally known additives, including carbon black, some plasticizers, anti-deterioration agents, vulcanizing agents, and vulcanization aids. The tension core 18 of the belt of System A was a twisted high-strength glass fiber yarn treated with an NBR / RFL treatment agent to counter the oil used in the oil environment.

The belt 3 is configured in the same manner as the belt 2 except that different tension core wires are used. Belt 3 is a carbon-glass hybrid core as described in US Pat. No. 7,682,274, ie a glass core replaced by a carbon fiber core yarn coated with a plurality of high-strength glass fiber yarns. Had had.

First test series of comparative system

The test equipment layout of the first test series included a standard tensioner and pulley with inertia matched to the standard pulley. The cam pulley was circular and machined with a 1.5 mm diameter pitch line differential (DPLD) and the crank pulley had a 1.45 mm DPLD. Both cam sprockets had inertia that matched standard pulleys, including standard VVT equipment. The tensioner 244 was a standard compact tensioner with a single eccentric center and symmetric damping, a mounting tension of about 500 N, and a smooth 62 mm diameter pulley. The width of the surface of the tensioner pulley was 24.5 mm, and the total width of the tensioner was about 36 mm. This was the widest element and therefore the overall width of the drive mechanism package of the comparative example was about 36 mm. The crank speed varied between 5000 and 6000 rpm, with 10 seconds of acceleration and 10 seconds of deceleration repeated.

Five drive systems were tested as shown in Table 2, and the results of the sixth system were estimated based on the other five. Only the belt was changed in this series. Belt 1 was used in three different widths, 18 mm, 12 mm, and 10 mm. Belt 2 was used in two different widths, 18 mm and 10 mm, and the results were estimated for a width of 12 mm. Two similar test devices were used. The system was run for a fixed time against Comparative Examples (“Comparison”) Systems 1 and 2 until the belts of the other systems were damaged. The results of the systems 3 and 6 show that the device 1 provides a slightly longer travel time than the device 2. Such changes in system testing are not unusual. These tests were used as reference values for the tests of the system of the present invention. Further testing was carried out with a goal of minimizing the size of the package of the system and with a 10 mm wide belt to avoid excessively long running times.

Comparative example and the second test series of the system of the present invention

The test layout of this second test series was similar to that of the first series, as shown in FIG. 35, but the test apparatus was the third apparatus, ie apparatus 3. The device 3 was different from the first and second devices, and the standard flywheel was included in the third device, but not in the devices 1 and 2. As shown in Table 3, the two drive systems were tested in device 3. The comparison system 7 included a standard tensioner and pulley whose inertia matched the standard pulley. The cam pulley was circular and machined with a 1.5 mm diameter pitch line differential (DPLD) and the crank pulley had a 1.45 mm DPLD. Both cam sprockets had inertia that matched standard pulleys, including standard VVT equipment. The tensioner 244 was a standard compact tensioner with a single eccentric center and symmetric damping, a mounting tension of about 500 N, and a smooth 62 mm diameter pulley. The width of the surface of the tensioner pulley was 24.5 mm, and the total width of the tensioner was about 36 mm. This was the widest element and therefore the overall width of the drive mechanism package of the comparative example was about 36 mm. In this test series, the crank speed was maintained at a constant value of 4750 rpm during each test (which is approximately the maximum resonant speed). A belt 2 with a width of 10 mm, an RPP116 tooth profile and a pitch of 9.525 mm was used for the comparison system 7.

The system 8 of the present invention has the same test layout, but the cam sprocket is oval and machined with a 1.5 mm diameter pitch line differential (DPLD). The oval intake cam sprocket 242 was phased at a position of 13 gcw with an eccentricity of 0.75 mm, i.e., displaced clockwise by 13 grooves from the engine reference value, as shown in FIG. It is a robe type. The oval exhaust cam sprocket 245 was phased at a position of 23.5 gcw with an eccentricity of 1.0 mm, i.e., angularly displaced clockwise by 23.5 grooves from the reference value, as shown in FIG. It is a 3-robe type. Both cam sprockets have the same inertia as standard pulleys.

The system 8 tensioner 244 was a compact tensioner according to the design described herein, with a mounting tension of approximately 500 N and a single eccentric center with a pulley having a diameter of 60 mm. Therefore, the tensioner had a pulley surface width of 14 mm and an overall width of about 16 mm. The tensioner was also described in US Pat. No. 6,609,988, which was granted, for example, to Liu and the like, and had an asymmetric damping of about 300 N.

The crank pulley also had a reduced surface width to fit the narrow belt and package width.

Belt 3 was used for System 8 with a width of 10 mm, an RPP 116 tooth profile, and a pitch of 9.525 mm.

The two systems ran until the belt was damaged. The results are shown in Table 3. The running time of Comparative Example System 7 at 690 hours is consistent with the result of Comparative Example System 6 at 777 hours. The result of the system 8 of the present invention is almost twice the running time of 1266 hours. This indicates that the combination of a narrow highly elastic belt with a circular sprocket and a special narrow tensioner significantly extends the belt life.

Table 3 contains the results of some timing errors, including before and after the test. The result of the improved timing error is believed to be a combination of effects, including the effect of higher belt elasticity, the effect of tensioner design, and the effect of oval sprocket. Table 3 shows some of the individual effects from other timings, some of which were performed on 12 mm wide belts as shown in Table 3. The end result is a very narrow (10 mm belt) with a maximum timing error of less than 1 ° at the peak peak value over the life of the belt.

In addition, Table 3 shows that the effective tension of the belt of the system is substantially reduced from 500N to 250N, the effect of which is largely due to the oval sprocket. The advantageous effect of this oval pulley is very important. It allows the mounting tension to be reduced without causing tooth skipping, which in turn reduces the wear rate of the lands.

Thus, the results of Systems 7 and 8 show very good overall system performance and belt life when the highly elastic belt is combined with oval sprockets and concentric tensioners. The result is considered to be further improved by optimizing the mounting tension. The oval sprocket significantly reduces the effective tension in the crank, which properly prolongs the belt life in the shear damage mode of the tooth. However, the observed damage mode was land wear (which is closely related to PV in the system (contact surface pressure x slip speed)), thus reducing the positive effect of oval sprockets on belt life. .. In other words, oval sprockets were used to reduce the effective tension, but if the mounting tension was also reduced (ie, lower PV between the belt and pulley), the belt life was further extended.

A belt width of 10 mm was selected to speed up the test. From these results, and other experience with the correlation between accelerated testing and more realistic applications, increasing the belt width to 14 mm results in the durability of System 8 in the test equipment for approximately 3500 hours. It is believed that this is expected to be related to the life of the vehicle of about 240,000 km. A belt width of 14 mm is applicable to the tensioners and pulleys mentioned above, allowing the overall width of the drive package to be 18 mm or less in size, which is about half the package width of a conventional drive mechanism with standard elements. be. It is also estimated that this drive system is easily about 30% lighter than a conventional drive system with standard elements.

Some additional features of the invention, particularly related to oval sprockets, are shown below.

Feature 1. The present invention relates to a synchronous belt drive system, wherein the synchronous belt drive system is a first oval having a toothed surface and at least one straight portion (16) disposed between two arc portions (14, 15). The sprocket (10) is provided, the arc portion has a fixed radius (R1, R2), the straight portion has a predetermined length, and the sprocket (300) has a toothed surface, and the sprocket has an endless toothed member (endless toothed member). Engaged with the first oval sprocket by 200), the first oval sprocket (10) has an angular displacement timing error between the sprocket and the first oval sprocket less than 1.5 degrees at the peak peak value. It has a size and a phase so as to be.

Feature 2. The synchronous belt drive system in feature 1 further comprises a second oval sprocket connected to a second rotational load, the second oval sprocket engages the endless toothed member, and the second oval sprocket is with the sprocket. It has a magnitude and phase such that the angular displacement timing error between the second oval sprockets is less than 1.5 degrees at the peak peak value.

Feature 3. In the synchronous belt drive system of feature 1, the angular displacement timing error between the sprocket and the first oval sprocket is less than 0.5 degrees at the peak peak value.

Feature 4. In the synchronous belt drive system of feature 3, the angular displacement timing error between the sprocket and the second oval sprocket is less than 0.5 degrees at the peak peak value.

Feature 5. In the synchronous belt drive system according to the feature 1, the width of the endless toothed member is 12 mm or more.

Feature 6. In the synchronous belt drive system of feature 1, the endless toothed member has elasticity in the range of about 630,000 N to about 902,000 N.

Feature 7. In the synchronous belt drive system of feature 1, the size is in the range of about 1.0 mm to 1.5 mm.

Feature 8. In the synchronous belt drive system of feature 1, the phase of the first oval sprocket is in the range of 9 to 25 grooves when rotated with respect to the reference point.

Feature 9. In the synchronous belt drive system in feature 8, the reference point is based on the 3 o'clock position.

Feature 10. In the synchronous belt drive system of feature 2, the phase of the second oval sprocket is in the range of 9 to 25 grooves when rotated with respect to the reference point.

Feature 11. In the synchronous belt drive system in feature 10, the reference point is relative to the 3 o'clock position.

Feature 12. In the synchronous belt drive system of feature 10, the phase of the first oval sprocket is in the range of 9 to 25 grooves when rotated with respect to the reference point.

Feature 13. In the synchronous belt drive system in feature 12, the reference point is relative to the 3 o'clock position.

Feature 14. In the synchronous belt drive system of feature 1, the sprocket is connected to the drive and the first oval sprocket is connected to the rotational load.

Feature 15. In the synchronous belt drive system of feature 14, the drive is the crankshaft of the engine.

Feature 16. In the synchronous belt drive system of feature 2, the first oval sprocket is connected to the exhaust camshaft.

Feature 17. In the synchronous belt drive system of feature 2, the second oval sprocket is connected to the intake camshaft.

Feature 18. The present invention also relates to a synchronous belt drive system, wherein the synchronous belt drive system comprises a first oval sprocket having a toothed surface and at least one straight portion arranged between two arc portions, wherein the arc portion is provided. The straight portion has a fixed radius, the straight portion has a predetermined length, and the sprocket has a toothed surface, the sprocket is engaged with the first oval sprocket by an endless toothed member, and the first oval sprocket is a sprocket. With a second oval sprocket connected to the second rotating load, with a magnitude and phase such that the angular displacement timing error between and the first oval sprocket is less than 1 degree at the peak peak value. The second oval sprocket engages the endless toothed member, and the second oval sprocket has an angular displacement timing error between the sprocket and the second oval sprocket that is less than 1.5 degrees at peak peak value. It has such a size and phase.

Feature 19. In the synchronous belt drive system of feature 18, the first oval sprocket is connected to the exhaust camshaft, the second oval sprocket is connected to the intake camshaft, and the sprocket is connected to the crankshaft of the engine.

Feature 20. In the synchronous belt drive system of feature 19, the angular displacement timing error between the sprocket and the first oval sprocket is less than 0.5 degrees at the peak peak value, and the angular displacement timing between the sprocket and the second oval sprocket. The error is less than 0.5 degrees at the peak peak value.

Some of the additional features of the invention related to tensioners are shown below.

Feature 1. The present invention relates to a tensioner, wherein the tensioner comprises a base having a cylindrical portion extending in the axial direction, the cylindrical portion having a radial outer peripheral surface and a receiving portion radially inside the radial outer peripheral surface. It is equipped with an eccentric arm that swingably engages with the outer peripheral surface and a torsion spring that is arranged in a receiving portion on the inner side in the radial direction. And.

Feature 2. In the tensioner in feature 1, the pulley is pivotally supported by a needle bearing.

Feature 3. In the tensioner in feature 1, the eccentric arm, the pulley, and the torsion spring are coaxially arranged so that none of the eccentric arm, the pulley, or the torsion spring is displaced in the axial direction along the axis AA from the other. ..

Feature 4. In the tensioner in feature 1, the eccentric arm is pivotally supported by the base in the bush.

Feature 5. The present invention also relates to a tensioner, wherein the tensioner has a base cylindrical portion having a radial outer peripheral surface and a radially inner receiving portion, an eccentric arm swingably engaged with the radial outer peripheral surface, and a radially inner receiving portion. The torsion spring is provided in the portion, and the torsion spring applies an urging force to the eccentric arm, and is provided with an elongated member arranged so as to engage with the eccentric arm and swing according to the rotation of the eccentric arm.

Feature 6. In the tensioner of feature 5, the eccentric arm and the torsion spring are coaxially arranged so that neither the eccentric arm nor the torsion spring is axially displaced along the axis AA from the other.

Feature 7. In the tensioner in feature 5, the eccentric arm is pivotally supported by the base in the bush.

Feature 8. In the tensioner in feature 5, the pulley is pivotally supported by the eccentric arm in the needle bearing.

Feature 9. The present invention also relates to a tensioner, which comprises a base having a cylindrical portion extending axially, the cylindrical portion having a radial outer peripheral surface and a radially inner receiving portion, which is swingable on the radial outer peripheral surface. It has an eccentric arm that engages with, a torsion spring that is located in the radially inner receiving part, the torsion spring applies urging force to the eccentric arm, and has a pulley that is pivotally supported by the eccentric arm. The pulley and the torsion spring are arranged coaxially so that none of the eccentric arm, the pulley, and the torsion spring is axially displaced along the axis AA from the eccentric arm, the pulley, or the torsion spring.

Feature 10. In the tensioner in feature 9, the base further has a fluid conduit, which allows the fluid to enter the bearing.

Feature 11. In the tensioner in feature 9, the pulley is pivotally supported by the bearing.

Feature 12. In the tensioner in feature 11, the bearing is a needle bearing.

Although the invention and its advantages have been described, it is understood that various modifications, substitutions and variations are possible without departing from the spirit and scope of the invention as defined by the appended claims. Should be. Furthermore, the scope of the invention is not intended to be limited to the specific embodiments of the processes, machines, products, compositions, means, methods, and steps described herein. As will be readily appreciated by those skilled in the art from the disclosure of the present invention, any existing or later developed will perform substantially the same functions as the corresponding embodiments described herein, or achieve substantially the same results. Processes, machines, products, compositions, means, methods, or steps to be made can be used in accordance with the present invention. Accordingly, the appended claims are intended to include the scope of such processes, machines, products, compositions, means, methods, or steps. The invention disclosed herein can be adequately implemented without elements not specifically disclosed herein.

Claims (11)

Synchronous belts (200) with tension cores with highly elastic fibers and
A drive sprocket and at least one driven sprocket, of which at least one is an oval sprocket (10).
A base comprising a tensioner, the tensioner having a radial outer peripheral surface and a receiving portion radially inside the radial outer peripheral surface, and an axially extending cylindrical portion, and the radial outer peripheral surface. It is provided with an eccentric arm that swingably engages, a torsion spring that is arranged in the radial inner receiving portion and applies urging force to the eccentric arm, and a pulley that is pivotally supported by the eccentric arm.
The eccentric arm, the pulley, and the torsion spring are coaxially arranged so as not to be displaced in the axial direction along the axial center of the cylindrical portion from any of the eccentric arm, the pulley, and the torsion spring. Being done
A synchronous belt drive system in which the pulley surrounds the entire eccentric arm. The oval sprocket (10) comprises a toothed surface and at least one straight line portion (16) disposed between the two arc portions (14, 15), the arc portion having a fixed radius (R1, R2). ), The synchronous belt drive system according to claim 1, wherein the straight portion has a predetermined length. The first aspect of claim 1, wherein the elliptical sprocket (10) has a size and a phase such that the angular displacement timing error between the driving sprocket and the driven sprocket is smaller than 0.5 degrees at the peak peak value. Synchronous belt drive system. The synchronous belt drive system according to claim 1, wherein the highly elastic fiber is one or more selected from the group consisting of fiberglass, PBO, aramid, and carbon fiber. The synchronous belt drive system according to claim 1, wherein the highly elastic fiber is a high-strength glass fiber. The synchronous belt drive system according to claim 1, wherein the highly elastic fiber is carbon fiber. The synchronous belt drive system according to claim 1, wherein the tension core wire is a hybrid core wire including carbon fiber and glass fiber. The synchronous belt drive system according to claim 1, wherein the width of the entire system is shorter than 20 mm. The synchronous belt drive system according to claim 1, wherein the width of the entire system is 18 mm or less. The synchronous belt drive system according to claim 1, wherein the width of the entire system is 16 mm or less. The synchronous belt drive system according to claim 1, wherein the width of the entire system is 14 mm or less.

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