WO2010071247A1

WO2010071247A1 - Multiple-clutch transmission, mct

Info

Publication number: WO2010071247A1
Application number: PCT/KR2008/007486
Authority: WO
Inventors: Myung Koo Kang
Original assignee: Myung Koo Kang
Priority date: 2008-12-16
Filing date: 2008-12-17
Publication date: 2010-06-24

Abstract

An improved multi-clutch transmission for transmitting a rotating force from an engine by changing a speed of the rotating force. The multi-clutch transmission includes an input shaft receiving the rotating force from the engine; first and second lay shafts arranged eccentric to an axis of rotation of the input shaft, wherein the first and second lay shafts are spaced apart from each other and parallel; first to third front gears for transmitting the rotating force of the input shaft to the first and second lay shafts; first and second clutches connected to the first and second lay shafts, respectively, to engage and disengage power; and an output shaft arranged between the first and second lay shafts to be parallel to the first and second lay shafts.

Description

Description MULTIPLE-CLUTCH TRANSMISSION, MCT

Technical Field

[1] The present invention relates to a transmission of a vehicle, and more particularly, to an improved multi-clutch transmission, which can quickly and stably perform gear shifting using a plurality of clutches and synchromesh mechanisms, advantageously utilize a space by reducing the volume of a gearbox using lay shafts and an output shaft, which can be freely arranged, reduce manufacturing costs due to a simple structure, and provide for convenient maintenance. Background Art

[2] The most important aspects in recent automobile technology are the improvement of fuel efficiency and the reduction in automobile pollution. While electric automobiles using fuel cells as a power source are getting the most attention, a great amount of time is still required until they are commercially viable. Therefore, the automobile industry is continuously endeavoring to cope with fuel and environmental problems as well as striving to ensure a better vehicle environment by raising fuel efficiency and reducing the amount of pollutants until the time when fuel cell vehicles can be commercialized.

[3] While most technical attempts have been concentrated on engines where fuel is mainly consumed, a variety of attempts are also being pursued in parts other than the engine. In particular, some recent transmission technologies have started to solve the problems of conventional transmissions thanks to new approaches.

[4] Types of transmissions include a Manual Transmission (MT), an Automatic

Transmission (AT), a Continuously Variable Transmission (CVT), an Automated Manual Transmission (AMT) (which has been introduced relatively recently) and a Dual-Clutch Transmission (DCT).

[5] The MT allows a driver to change gears by directly operating a gear lever, and the

AT is a transmission system that includes a torque converter, planetary gears, a hydraulic clutch and a brake to automatically perform gear shifting. The CVT is a type of AT system, in which pulleys each having a side plate with a variable interval are attached to power input and output shafts, respectively, and are connected to each other by a steel belt or chain. More attention is being paid recently to the CVT as an alternative to the AT since the CVT has similar drive characteristics but better fuel efficiency compared to the AT. The CVT is widening its scope of application to include large vehicles.

[6] The AMT is a transmission system that has the same gear construction and gear lever as those of the MT but which does not have a clutch pedal. The AMT was introduced in Korea using the term "semi-auto transmission". The AMT has merits such as inexpensive manufacturing costs compared to other types of transmissions and economic competitiveness at the level of the MT. However, interest in the AMT seems to have flagged these days.

[7] The DCT can be referred to as an advanced form of the AMT. Lately, the DCT has been in the spotlight since it has merits such as the high economic competitiveness of the MT, less power loss, high efficiency and sporty driving characteristics due to quick gear shifting. The DCT has been used in mass-produced vehicles since 2003, and at present, is most widely used in a Direct Shift Gearbox (DSG) of Volkswagen based on the Dualtronic technology of Borg Warner. While the DCT is similar to the MT in that gear shifting is enabled because of the assistance of an electronic hydraulic device using the structure of the MT as a basis therefor, it has great structural differences from the MT. In the MT, a gear shifting process generally includes disconnecting power transmission by disengaging a clutch in a state where gears are meshed with each other, operating a gear lever to an intended shift stage, and then connecting power transmission by engaging the clutch. Here, a procedure of regulating an accelerator is used in order to set the different numbers of rotation of two gears to be the same. The foregoing process is carried out by the hand and foot of a driver in the MT, and by the electronic hydraulic device in the AMT. Since one clutch is used, engaging and disengaging the clutch and selecting a gear should be carried out in sequence. Thus, shift speed has an upper limit and a shift shock may take place when the driver awkwardly operates the clutch.

[8] Table 1 below reports minimum shift times of vehicles using AMT.

[9]

[10] Table 1

[Table 1] [Table ]

[H] [12] FIG. 1 is a configuration view illustrating a conventional dual-clutch transmission. The Dual-Clutch Transmission (DCT) is different from the MT or AMT in that the DCT uses two clutches (clutches 1 and 2) as expected from its name. Describing the DCT with respect to a Volkswagen DSG (PDK) shown in the figure, clutches 1 and 2 transmit power through odd drive gears Dl, D3 and D5 and through even drive gears D2, D4 and D6, respectively. When one gear is selected, a gear for the shift stage adjacent to the shift stage of the selected gear is rotating in a pre-selected state, connected to an input shaft 1 or an input shaft 2. For example, in case that the first drive gear Dl is selected, when the clutch 1 for the odd drive gears is engaged to transmit power through the first drive gear Dl, the second drive gear D2 is also rotating but power is not transmitted since the clutch 2 for the even drive gears is disengaged. When the first drive gear is changed (or shifted) into the second drive gear, the clutch for the odd drive gears is disengaged and at the same time the clutch for the even drive gears is engaged to transmit power through the second drive gear. In this manner, whenever gear shifting is performed, power transmission is simultaneously connected and disconnected by the clutches, thereby performing gear shifting very quickly. Therefore, the DCT has characteristics such as a very small shift shock since a shock associated with engaging and disengaging the clutches is removed.

[13] The shift time of the DCT is approximately in the range from 8 to 10 ms, which is significantly improved over the foregoing AMT. In addition, the DCT has excellent fuel efficiency since it has less power loss than the MT. Owing to these characteristics, at present, worldwide vehicle makers and transmission providers are endeavoring and intensively competing to develop an improved DCT.

[14] In spite of these advantages, the conventional DCT has several drawbacks. As shown in FIG. 1, the conventional DCT has a very complicated structure, in which all components are very closely packed in a limited space. Thus, fabrication is difficult and it is expensive. In addition, the first input shaft 1 rotates inside the hollow funnel- shaped second input shaft 2, with the center of rotation of the first input shaft 1 concentric with that of the second input shaft 2. This, however, is an unstable structure. For example, when impurities intrude between the two input shafts, which are rotating at high speed, the hollow second input shaft may be damaged. This damage in the rotating input shaft can lead to a severe accident. Furthermore, the dual-rotary shaft structure is not easy to design or fabricate, and high precision is required since respective components are very closely arranged. Therefore, the problems are that manufacturing costs are very expensive and repair is very difficult when it does break. In addition, the second input shaft has insufficient durability since it is a hollow pipe. Furthermore, even though the DCT illustrated in the figure has a dual shaft structure with the center of rotation of the first input shaft concentric with that of the second input shaft, ability to reduce the volume of the DCT is limited since the two output shafts are arranged side by side as upper and lower output shafts.

[15] Moreover, the dual clutches 1 and 2 are implemented with a small multi-plate wet clutch to reduce the volume of the transmission. However, the wet clutch inevitably causes power loss since it uses cooling oil, and the small multiple plates of the clutch cannot sufficiently withstand high level torque.

[16] FIGS. 2 and 3 are front cross-sectional view illustrating the construction of the input side of another conventional DCT. The DCT is disclosed in Korean Patent Application Publication No. 10-2007-0104657, titled "POWER FLOW CONFIGURATION FOR DUAL CLUTCH TRANSMISSION MECHANISM." This document discloses a chain, gears and combinations thereof as an input side mechanism. FIG. 2 illustrates a case in which a chain 40 and sprockets 28, 42 and 44 are used as input-side power transmitting means, and FIG. 3 illustrates a case in which gears 1210, 1216 and 1218 and two idling gears 1212 and 1214 are used as input-side power transmitting means. The DCT using a chain as means for transmitting input torque from an engine has several problems unlike a timing belt using a chain. When the chain is used as shown in FIG. 2, power loss rate is high since the chain cannot reliably transmit power, unlike gears. In addition, the endurance of the chain, when it is used for a long time, is significantly poor compared to gears. In FIG. 3, the two idling gears 1212 and 1214 are used to transmit torque from the engine to the first and second clutch drive gears 1216 and 1218, respectively. However, this structure uses an excessively large number of gears to transmit power from the engine to first and second lay shafts. Thus, the problems are high power loss and expensive manufacturing and maintenance costs.

[17] In addition, like the earlier-described conventional art, this structure cannot escape the problem of power loss since it uses multi-plate wet clutches. The small multiple plates of the clutches cannot sufficiently withstand higher torque. Furthermore, this structure makes it difficult to reduce the volume of the gearbox by changing the position of the clutches.

[18] Even if the above-identified patent application discloses a combination of gears and a chain as input-side power transmitting means, this structure also has the foregoing problems related to the chain and the idling gears.

[19] The DCT of the above-identified patent application has a critical drawback of large size since it fails to reduce the volume of the gearbox, which is the problem of the most recent transmissions waiting to be solved. Therefore, there are still many problems to be solved before commercialization.

[20] Accordingly, there are strong demands for the development of a novel transmission that can compensate for most of the drawbacks while utilizing the advantages of the conventional DCT. That is, what is required is the development of a novel transmission that can more reliably transmit power, facilitate reduction of the volume of a gearbox due to high flexibility in the arrangement and combination of clutches, and facilitate designing and controlling the gear ratio of shift gears. Disclosure of Invention

Technical Problem

[21] The present invention provides a multi-clutch transmission (MCT), which is designed to improve the arrangement and structure of clutches, an input shaft, and an output shaft so that gear shifting is performed quickly and stably, manufacturing and maintenance costs and weight can be easily reduced due to the simple structure.

[22] An object of the present invention is to transmit torque from an engine to first and second lay shafts through direct contact using three front gears, thereby simplifying the structure of the transmission and raising the efficiency of power transmission.

[23] Another object of the present invention is to minimize the volume of the gearbox and thereby improve fuel efficiency and optimize the design of an engine room by changing the combination and arrangement of the clutches or adopting a gearbox volume correction gear.

[24] A further object of the present invention is to facilitate designing and controlling the gear ratio and using a common gear by changing the position of the axis of the output shaft in vertical and horizontal directions with respect to the axis of the input shaft or by using a gear ratio control gear. Technical Solution

[25] According to an aspect of the present invention, there is provided a multi-clutch transmission for transmitting a rotating force from an engine by changing a speed of the rotating force. The multi-clutch transmission may include an input shaft 210 receiving the rotating force from the engine 200; first and second lay shafts 110 and 120 arranged eccentric to an axis of rotation of the input shaft, wherein the first and second lay shafts are spaced apart from each other and parallel; first to third front gears Fl, F2 and F3 for transmitting the rotating force of the input shaft to the first and second lay shafts; first and second clutches Cl and C2 connected to the first and second lay shafts, respectively, to engage and disengage power; and an output shaft 130 arranged between the first and second lay shafts to be parallel to the first and second lay shafts.

[26] The first front gear Fl may be axially mounted on the input shaft 210, the second front gear F2 may be axially mounted on the first lay shaft 110, and the third front gear F3 may be axially mounted on the second lay shaft 120, wherein the second and third front gears have an equal number of teeth and an equal diameter.

[27] Drive gears of an odd shift stage may be arranged on the first lay shaft, and drive gears of an even shift stage are arranged on the second lay shaft.

[28] The multi-clutch transmission may further include a plurality of output gears arranged on the output shaft 130, wherein the output gears are common gears each of which is meshed with a corresponding one of the drive gears of the odd shift stage and a corresponding one of the drive gears of the even shift stage.

[29] At least one of the common gears may include a gear ratio control gear that is provided integratedly with the output gear.

[30] The multi-clutch transmission may further include a plurality of output gears arranged on the output shaft 130, wherein each of the output gears is meshed with a corresponding one of the drive gears of the odd or even shift stage.

[31] The second clutch C2 may be positioned on a plane including the first clutch Cl or on a plane different from that including the first clutch Cl.

[32] The first and second clutches Cl and C2 may have different diameters.

[33] The multi-clutch transmission may further include a main clutch MC arranged between the engine 200 and the first to third front gears Fl, F2 and F3 and connected to the input shaft.

[34] The output shaft 130 may be eccentric to the axis of the input shaft 210. Here, the output shaft 130 can be displaced in parallel so as not to be spaced at the same distance from the first and second lay shafts 110 and 120.

[35] The multi-clutch transmission may further include gearbox volume correction gears CGl, CG2, CG3 and CG4 interposed between the first clutch Cl and the first lay shaft 110 and between the second clutch C2 and the second lay shaft 120, thus reducing the axial distance between the first and second lay shafts.

[36] The multi-clutch transmission may further include a plurality of synchromesh mechanisms arranged on the first lay shaft 110 and the second lay shaft 120, wherein each of the synchromesh mechanisms maintains a pre-selected state coupled with a drive gear of a next shift stage when a selected drive gear of an odd or even shift stage is outputting the rotating force.

[37] The input shaft 210 and the output shaft 130 may be positioned or not on a plane including the first and second lay shafts 110 and 120.

Advantageous Effects

[38] The multi-clutch transmission of the present invention has a simple structure, which leads to inexpensive manufacturing costs, easy maintenance and excellent durability due to rare trouble.

[39] Since the two lay shafts and the output shaft do not need to be arranged concentrically with the input shaft that directly receives a rotating force from an engine and can be freely arranged on different planes, the position and size of a gearbox can be freely adjusted according to the structure, drive mode, output and use of a vehicle. These advantageous features allow the gearbox to be designed using a compact structure by reducing its size, thereby improving fuel efficiency.

[40] In addition, the gear ratio of the transmission gear can be freely designed since the input shaft, the two lay shafts and the output shaft can be combined by freely changing positions.

[41] Furthermore, there are advantages such as a wide application range due to a structure capable of employing conventional multi-plate wet clutches and large dry clutches, and the clutches can be freely constructed in the form of dual or triple clutches.

[42] The effects of the present invention are described more fully in comparison with conventional dual-clutch transmissions.

[43] First, the present invention has advantages of less power loss and reliable power transmission since rotation and torque from an engine are transmitted to first and second lay shafts through a front gear axially mounted on an input shaft and two front gears meshed with the front gear and in direct contact therewith, thus simplifying the route of power transmission.

[44] Second, unlike conventional dual clutch transmissions that control the gear ratio based on a value obtained by multiplying an input-side gear ratio (i.e., upstream gear ratio) with an output-side gear ratio (i.e., downstream gear ratio), the present invention designs and controls the gear ratio by changing the position of the axis of the output shaft in vertical and/or horizontal directions or by using a separate gear ratio control gear, therefore the gear ratio can be designed and changed easily and it is not required to design odd and even clutch-driving sprockets or gears with different sizes. In addition, when the gear ratio control gear is employed, all output gears on the output shaft can be constructed as a common gear unlike conventional transmissions. Therefore, the length of the output shaft can be reduced to be shorter than that of conventional transmissions, and thereby the entire length of the transmission of the invention can be significantly reduced.

[45] Third, in the present invention, the size of the clutch can be changed without increasing the volume of the gearbox, and, when necessary, inconsistencies between the area of the clutches and the size of the gearbox can be removed by arranging the clutches on different planes. In addition, the separate gearbox volume correction gear can significantly reduce the width of the gearbox to optimize the design of an engine room so that other adjacent components can be freely designed and arranged.

[46] Fourth, in the present invention, a main clutch is provided in front of the first and second clutches to reduce the sizes required of the first and second clutches. This makes it possible to provide a main clutch having a size sufficient to manage a large amount of torque, which is necessary for a low gear of a large vehicle or the like. As a result, the width of the gearbox can also be reduced.

[47] Fifth, in the present invention, the diameter or number of teeth of the second front gear which drive the first lay shaft is not required to be different from that of the third front gear for driving the second lay shaft since the gear ratio of a shift stage can be controlled using only the displacement of the output shaft or the gear ratio control gear. Here, gears of the same dimensions can be used so that the second and third front gears can have the same diameter or number of teeth. Accordingly, manufacturing and maintenance costs of the gears can be reduced.

[48] Sixth, since the second and third front gears are arranged in front of the first and second clutches, the multi-clutch transmission of the present invention has an advantage of a simpler and more stable structure compared to the conventional transmission in which a clutch-driving sprocket or clutch-driving gears are arranged behind a multi-plate wet clutch. Brief Description of Drawings

[49] FIG. 1 is a configuration view illustrating a conventional DCT (PDK);

[50] FIGS. 2 and 3 are front cross-sectional views illustrating the construction of the input side of another conventional DCT;

[51] FIG. 4 is a schematic configuration view illustrating a multi-clutch transmission according to the present invention; [52] FIGS. 5 to 12 are block diagrams illustrating power transmission procedures from a first-speed mode to a reverse mode in the multi-clutch transmission according to the present invention;

[53] FIG. 13 is a schematic configuration view illustrating a multi-clutch transmission according to another embodiment of the present invention, in which the position of a shift stage is changed;

[54] FIG. 14 is a schematic configuration view illustrating an example of the present invention in which a gear ratio is controlled by changing the position of an output shaft;

[55] FIGS. 15 to 18 are schematic configuration views illustrating a variety of combinations including first and second clutches according to the present invention;

[56] FIG. 19 is a schematic configuration view illustrating a clutch according to another embodiment of the present invention;

[57] FIG. 20 is a schematic configuration view illustrating a multi-clutch transmission without a common gear;

[58] FIG. 21 is a schematic configuration view illustrating an example of an output gear of the present invention, with which a gear ratio control gear is provided;

[59] FIG. 22 is a schematic configuration view illustrating a multi-clutch transmission in which the width of a gearbox is reduced using a gearbox volume correction gear of the present invention;

[60] FIGS. 23 to 31 are schematic configuration views illustrating a variety of arrangements of an input shaft, an output shaft and first and second lay shafts of the present invention; and

[61] FIG. 32 is a schematic plan view for comparing the sizes of gearboxes using conventional transmissions with that of a gearbox using a multi-clutch transmission of the present invention.

[62] <Major Reference Numerals of the Drawings>

[63] 110: first lay shaft 120: second lay shaft

[64] 130: output shaft 210: input shaft

[65] 200: engine

[66] Fl, F2, F3: first, second and third front gears

[67] Cl, C2: first and second clutches

[68] Dl to D7: drive gears R: reverse gear

[69] Gl to G7: output gears (driven gears)

[70] G6: gear ratio control gear (driven gear)

[71] CGl to CG4: gearbox volume correction gear

[72] Sl to S4: synchromesh mechanism

[73] DF: differential gear IG: idling gear [74] MC: main clutch

Best Mode for Carrying out the Invention

[75] The construction and operation of the present invention will now be described more fully hereinafter with reference to the accompanying drawings.

[76] FIG. 4 is a schematic configuration view illustrating a multi-clutch transmission according to the present invention. The Multi-Clutch Transmission (MCT) according to the present invention can basically include two or three clutches. Referring to FIG. 4, the multi-clutch transmission is illustrated as including two clutches Cl and C2. As shown in the figure, the multi-clutch transmission according to one embodiment of the present invention includes an input shaft (engine shaft) 210, first to third front gears Fl to F3, first and second lay shafts 110 and 120, the first and second clutches Cl and C2, first to seventh drive gears Dl to D7, a reverse gear R, output gears (driven gears) Gl to G4 and an output shaft 130.

[77] The input shaft 210 is also referred to as an engine shaft and it receives a rotating force from an engine 200.

[78] The first and second lay shafts 110 and 120 are arranged to be eccentric from the axial center (i.e., axis) of the input shaft 210 but to be parallel to each other.

[79] A single output shaft 130 is provided between the first and second parallel lay shafts

110 and 120 and is parallel to the respective lay shafts. The output shaft 130 can be provided with the center of rotation (i.e., axis) being coaxial with or eccentric from the input shaft 210.

[80] The first, third, fifth and seventh drive gears Dl, D3, D5 and D7 related with odd shift stages are arranged in sequence on the first lay shaft 110, and the second, fourth and sixth drive gears D2, D4 and D6 and the reverse gear R related with even shift stages and reverse shift stage respectively are arranged in sequence on the second lay shaft 120.

[81] The first and second clutches Cl and C2 selectively transmit the rotating force

(including torque and speed) of the input shaft 210, received from the engine, to the first lay shaft 110 or the second lay shaft 120. When the first clutch Cl is activated, the rotating force of the input shaft 210 is transmitted to the first lay shaft 110, and the second clutch C2 remains in an inactivated state. Conversely, when the second clutch C2 is activated, the rotating force of the input shaft 210 is transmitted to the second lay shaft 120, and the first clutch Cl is in an inactivated state.

[82] The output shaft 130 is meshed with a differential gear DF to transmit the rotating force of the input shaft 210 finally to the wheels through the transmission.

[83] On the shaft between the drive gears, there are slidably coupled synchromesh mechanisms Sl to S4, which are selectively coupled with drive gears to transmit power. The synchromesh mechanisms Sl to S4 are connected to a separate actuator (not shown) and are controlled by an electronic control unit. The first and second clutches Cl and C2 are basically implemented with a wet multi-plate clutch, but can be implemented with a dry clutch when necessary.

[84] FIGS. 5 to 12 are block diagrams illustrating power transmission procedures from a first-speed mode to a reverse mode in the multi-clutch transmission according to the present invention.

[85] In first-speed mode, as shown in FIG. 5, the rotating force from the engine 200 is transmitted to the first lay shaft 110 through the input shaft 210. The first clutch Cl is engaged to transmit power, but the second clutch C2 is disengaged. The first synchromesh mechanism Sl is coupled with the first drive gear Dl, allowing the rotating force to be transmitted to the first drive gear Dl. Since the first drive gear Dl is meshed with the first output gear Gl which is a driven gear, the rotating force from the engine is transmitted to the differential gear DF through the first output gear (driven gear) Gl and the output shaft 130. Here, the second synchromesh mechanism S2 is previously selected to prepare for the next gear shifting and remains coupled with the second drive gear D2. This state is referred to as a pre-selected state. However, even if the second drive gear D2 is pre-selected, the rotating force of the input shaft 210 is not transmitted to the second drive gear D2 since the second clutch C2 connected to the second lay shaft 120 is disengaged to idle the second lay shaft 120. The thick solid line in the figure indicates a route through which power is transmitted.

[86] In second-speed mode, as shown in FIG. 6, the first clutch Cl is disengaged and the second clutch C2 is engaged. At this time, since the second drive gear D2 is already pre-selected, the rotating force is immediately transmitted to the second drive gear D2 as soon as the second clutch C2 is engaged. The second drive gear D2 in turn transmits the rotating force to the first output gear (driven gear) Gl to rotate the output shaft 130. As in the foregoing first-speed mode, the first synchromesh mechanism Sl is coupled with the third drive gear D3 for the next gear shifting.

[87] The operations from third-speed mode to reverse mode are performed in such a fashion that the first and second clutches Cl and C2 are alternately engaged or disengaged to transmit power as shown in FIGS. 7 to 12. In addition, similarly to the foregoing modes, a drive gear of one shift stage corresponding to the next stage is previously selected and waits in a pre-selected state. In this manner, gear shifting can be quickly performed from a first-speed gear to a seventh-speed gear. In reverse mode, as shown in FIG. 12, the fourth synchromesh mechanism S4 is coupled with the reverse gear R to transmit the rotating force to the fourth output gear (driven gear) G4 through an idling gear IG.

[88] The foregoing gear shifting according to the present invention can be manually or au- tomatically carried out using a shift button, a shift paddle or a general shift lever. The multi-clutch transmission basically has the merits of a manual gear but does not need a clutch pedal since the gear shifting is automatically performed by an electronic hydraulic actuator. Since the multi-clutch transmission of the present invention is constructed to directly operate a pre-selected drive gear corresponding to the next stage without a time difference by alternately operating the two clutches Cl and C2, it is not required to perform a gear- shifting operation by stepping on and off the clutch pedal as in a manual transmission. The electronic control unit of a vehicle selects a drive gear corresponding to the next stage by determining that a driver will select the next stage drive gear based on data read from the position of a throttle and the count of revolutions. For example, in the state where the second drive gear D2 is currently operating, pre-selecting the third drive gear D3 does not influence the operating second drive gear D2 since the clutch Cl related with the odd shift stage is disengaged. When the driver operates a gear shift pedal, the electronic control unit disengages the clutch C2 related with the even shift stage and at the same time sends a signal instructing the clutch Cl related with the odd shift stage to get connected. In this manner, the second drive gear is promptly changed into the third drive gear without any delay. Accordingly, the state in which power is not transmitted at all does not exist unlike the manual transmission, and thus gear shifting is performed very quickly and smoothly without braking. When the multi-clutch transmission of the present invention is used, gear shift time can be reduced to 8 ms or less. [89] Unlike a conventional PDK dual-clutch transmission, in this embodiment of the present invention, the two lay shafts 110 and 120 are arranged eccentric from the input shaft 210 which directly receives power from the engine. The center of rotation (axis) of the two lay shafts 110 and 120 is not consistent with that of the input shaft 210. In addition, the first lay shaft 110 and the second lay shaft 120 are spaced apart from each other at a predetermined distance and arranged to be parallel to each other unlike the conventional structure in which first and second lay shafts share the center of rotation (axis) with the first lay shaft housed inside the hollow second lay shaft. This structure has merits such as a simple structure, easier maintenance and an inexpensive manufacturing cost compared to the conventional PDK dual-clutch transmission. Owing to such structural differences, unlike the conventional PDK dual-clutch transmission, the present invention makes it possible to arrange the first to third front gears Fl to F3 in front of the first and second clutches Cl and C2 to directly transmit the rotating force from the input shaft 210 to the first and second lay shafts 110 and 120. In the present invention, the output shaft 130 is singular in number, and the center of rotation (axis) of the output shaft 130 can be arranged to be consistent or inconsistent with that of the input shaft 210. The transmission can be optimally constructed according to a driving mode (e.g., front or rear wheel drive), output power, size and a usage of a vehicle by suitably arranging the mutual positions of the front gears Fl to F3, the input shafts 110 and 120 and the output shaft 130. Related details will be described later on.

[90] In this embodiment, the first output gear Gl is shared by the first drive gear Dl and the second drive gear D2. Thus, both the first and second drive gears Dl and D2 are meshed with the first output gear Gl to transmit the rotating force to the output shaft 130. This structure can reduce the number of output gears on the output shaft 130 to decrease the length of the output shaft 130 and thus properly reduce the size of the transmission. Due to such a merit, a high-performance transmission can be advantageously applied to a small automobile. Further, this merit can lead to effects of increasing fuel efficiency while preventing excessive fuel consumption due to an increased volume of the transmission.

[91] While this embodiment has been described with respect to the transmission having a total of eight drive gears including the reverse gear, it should not be understood that the number of the drive gears is limited. Rather, the number of the drive gears can be altered without limitation.

[92] FIG. 13 is a schematic configuration view illustrating a multi-clutch transmission according to another embodiment of the present invention, in which the position of a shift stage is changed. In general, greatest torque is applied to the first drive gear Dl of the components of the transmission, and the next greatest torque is applied to the reverse gear R and then to the second drive gear D2. Thus, when the reverse gear R2 is required to be subjected to greater torque than the second drive gear D2, as shown in the figure, it can be preferable that the first, third and fifth drive gears Dl, D3 and D5 and the reverse gear R are arranged on the first lay shaft 110 whereas the second, fourth, sixth and seventh drive gears D2, D4, D6 and D7 are arranged on the second lay shaft 120. The size of the first clutch Cl connected to the first drive gear Dl and the reverse gear R can be preferably greater than that of the second clutch C2 connected to the second drive gear D2. In this case, an increasing volume of the gearbox due to the increased size of the first clutch Cl can be prevented using a gearbox volume correction gear, which will be described later on, or by arranging the clutches on different planes.

[93] FIG. 14 is a schematic configuration view illustrating an example of the present invention in which a gear ratio is controlled by changing the position of an output shaft. In the conventional dual-clutch transmission shown in FIGS. 2 and 3, an effective gear ratio is determined by multiplying an upstream gear ratio with a downstream gear ratio, in which the upstream gear ratio is a gear ratio between the engine input sprocket 28 for transmitting power from the engine and the odd and even clutch driving sprockets 42 and 44 or between the gear 1210 and the gears 1216 and 1218, and the downstream gear ratio is a gear ratio between the pinion on the lay shaft and a gear on the output shaft. Thus, it is difficult to set or adjust the effective gear ratio of a shift stage. It is also very uneconomic since the size of the sprocket and the gears of the transmission are required to be entirely changed. However, in the present invention, an intended gear ratio can be set or controlled by adjusting a displacement d of the output shaft 130 as shown in the figures. When a gear ratio control gear (see G6 in FIG. 21), which will be described later, is provided to a drive gear (see G5 in FIG. 21) on the output shaft 130, the gear ratio of the shift stage can be variously controlled even if a common gear is used.

[94] FIGS. 15 to 18 are schematic configuration views illustrating a variety of combinations of first and second clutches according to the present invention. Since the conventional dual-clutch transmission (PDK) is constructed with two clutches and two output shafts, a wet multi-plate clutch of a small diameter is employed in order to prevent the volume of the transmission from increasing. As drawbacks compared to a typical dry clutch, the wet multi-plate clutch has a large amount of power loss and can rarely withstand a large amount of torque. However, the present invention minimizes the volume of the transmission by spacing the two lay shafts 110 and 120 apart from each other and placing the output shaft 130 between the lay shafts 110 and 120, and thus dry clutches which are relatively bulky and have good performance on a great amount of torque can be applied. Of course, it will be apparent to those skilled in the art that this does not exclude the application of the wet multi-plate clutch as clutching means of the present invention. In other words, the wet clutch is basically used as the clutch of the present invention, but the use of the dry clutch is not excluded. When the wet clutch is used, unlike the conventional art, the two wet multi-plate clutches Cl and C2 can be arranged in such a manner that the two clutches Cl and C2 are positioned on different planes such that predetermined areas of clutch plates overlap each other to reduce the space occupied by the clutches. As a result, the number of necessary clutch plates can be reduced to decrease the length of the transmission or a gearbox.

[95] The number of shift stages and the gear ratio are changed according to engine power, torque, vehicle weight, purpose of use, etc., which are determined when designing a vehicle. FIG. 15 illustrates an example in which the two clutches Cl and C2 are designed with the same size, FIG. 16 illustrates an example in which the first clutch Cl coupled with a first-speed gear is designed to be greater than the second clutch C2. The greater clutch can more securely transmit power since it can withstand a greater amount of torque. However, since the size of the gearbox is limited, it is important to properly adjust the size of the clutch to find an optimal condition. The size of the first clutch Cl is important since the first-speed gear has the greatest static friction force when a vehicle starts on an inclined road, etc. Thus, the size of the clutch is determined considering all situations of the vehicle such that the clutch can be securely engaged and disengaged without slipping. While the two clutches Cl and C2 can be preferably designed to be of the same size, the first clutch Cl can be designed to be greater than the second clutch C2 as shown in FIG. 16. Then, the width and the resultant volume of the gearbox of this construction are greater than those of FIG. 15. In this case, as shown in FIG. 17, the relatively greater first clutch Cl and the relatively smaller second clutch C2 can be arranged in different planes such that predetermined areas of the clutch plates overlap each other. Thereby, a clutch of a greater size capable of withstanding a sufficient amount of torque can be employed without increasing the width of the gearbox.

[96] FIG. 18 illustrates a design in which both the clutches Cl and C2 are designed to be of a predetermined size or more. This design can be used when a great amount of torque is applied to both first and second speed gears. Unlike FIGS. 15 and 16, FIGS. 17 and 18 show the two clutches Cl and C2 arranged in different planes.

[97] As described above, both the dry and wet clutches can be applied to the transmission of the present invention. Even in the case where the dry clutches are used, the two clutches can be advantageously arranged in a variety of sizes and in a variety of positions according to the intended use.

[98] FIG. 19 is a schematic configuration view illustrating a clutch according to another embodiment of the present invention. In this embodiment, unlike the foregoing embodiments, a separate main clutch MC is added between the engine 200 and the first and second clutches Cl and C2. This construction is made on the basis that a clutch of a large size is required in the case where a vehicle starts on a road with a steep inclination. While the two clutches Cl and C2 are designed with different sizes in the foregoing embodiments shown in FIGS. 15 to 18, the main clutch MC is employed in this embodiment. When a large- sized main clutch MC is provided to engage or disengage a power transmission when a stopped vehicle starts or moves backwards, the remaining two clutches Cl and C2 can be designed in a relatively smaller size. Accordingly, the vertical width of the gearbox can be significantly decreased, thereby reducing the size of the gearbox. Here, the main clutch MC can be basically a wet multi-plate clutch but a dry clutch is not excluded. A torque converter of an automatic transmission can be used as the main clutch MC.

[99] FIG. 20 is a schematic configuration view illustrating a multi-clutch transmission without a common gear. In the illustrated construction, adjacent drive gears of odd and even shift stages do not share one output gear. Rather, each of the drive gears is meshed with a separate output gear. Accordingly, a total of eight output gears Gl to G7 and GR are arranged in sequence on the output shaft 130. In this construction, as a drawback, the increased length of the output shaft 130 increases the size and the resultant volume of the gearbox. However, the problem of the increased length of the gearbox can be overcome using a common gear, which will be described later.

[100] FIG. 21 is a schematic configuration view illustrating an example of an output gear of the present invention. The problem of the increasing size of the gearbox occurring in the embodiment shown in FIG. 20 can be overcome by constructing a gear combination as shown in FIG. 21. In addition, when the gear ratio of a shift stage is controlled by changing the relative position of the output shaft 130 with respect to the axis of the input shaft 210, the gear ratio control gear G6 shown in the figure can be used to facilitate designing and modifying the gear ratio. For example, the fifth output gear G5 and the adjacent sixth output gear G6 shown in FIG. 20 are provided in one gear block such that the respective output gears G5 and G6 are meshed with the fifth drive gear D5 and the sixth drive gear D6, respectively. Arranging the output gears on the output shaft 130 in this manner can greatly reduce the length of the output shaft 130 and thus decrease the size of the gearbox, thereby providing a compact transmission construction. Here, the fifth output gear G5 and the sixth output gear G6 are provided in one body, making a common gear that is meshed with both the fifth drive gear D5 and the sixth drive gear D6. Here, the sixth output gear G6 is defined as the gear ratio control gear. The common gear equipped with the gear ratio control gear is not limited to the case shown in FIG. 20 in which the fifth output gear G5 and the sixth output gear G6 are provided, but can also be applied to the foregoing cases shown in FIGS. 4 to 14 in which all the output gears are common gears. Furthermore, the gear ratio control gear can be employed when it is difficult to set the designed gear ratio of a shift stage in an attempt to control the gear ratio by changing the position of the output shaft 130. Then, the output gears of all shift stages can be provided as common gears to thereby greatly reduce the length of the gearbox.

[101] The transmission of the present invention can be easily designed to have an optimum gear ratio according to type, intended use and power levels of vehicles since the two lay shafts 110 and 120, the output shaft 130 and the input shaft 210 can be eccentrically arranged.

[102] FIG. 22 is a schematic configuration view illustrating a multi-clutch transmission in which the width of a gearbox is reduced using a gearbox volume correction gear of the present invention. In this embodiment, gearbox volume correction gears CGl to CG4 are employed in order to more greatly reduce the volume of the gearbox. The first and second lay shafts 110 and 120 are not directly connected to the first and second clutches Cl and C2, respectively. Instead, the first lay shaft 110 is connected to the first clutch Cl through the intermediate gearbox volume correction gears CGl and CG2 to transmit power, and the second lay shaft 120 is connected to the second clutch C2 through the intermediate gearbox volume correction gears CG3 and CG4 to transmit power. It should be understood that the width of the gearbox can be greatly reduced using the gearbox volume correction gears CGl to CG4. The availability of the gearbox volume correction gears will be apparent when a width dl of the gearbox before the use of the gearbox volume correction gears is compared with a width d2 of the gearbox after the use of the gearbox volume correction gears.

[103] FIGS. 23 to 31 are schematic configuration views illustrating a variety of arrangements of an input shaft, an output shaft and first and second lay shafts of the present invention. The figures schematically illustrate configurations viewed from the front of the transmission.

[104] In the present invention, the first and second lay shafts 110 and 120 of the transmission are not required to be arranged concentrically with the input shaft 210, and the output shaft 130 is not required to be arranged concentrically with the input shaft 210, either. Accordingly, the transmission can be constructed in various combinations as illustrated in the figures. Due to these merits, the transmission of the invention can advantageously design the gear ratio to be optimized to a vehicle, and the shape of the gearbox can be freely designed according to the size and power of the vehicle.

[105] Referring to FIGS. 23 to 25, axes of rotation of three front gears Fl to F3 are arranged on the plane, and an input shaft (engine shaft) ES and an output shaft (drive shaft) DS are eccentrically arranged. The second front gear F2 and an odd drive gear DO are arranged on a first lay shaft LSI, and the third front gear F3 and an even drive gear DE are arranged on a second lay shaft LS2. Here, the input shaft ES, the output shaft DS and the first and second lay shafts LS 1 and LS2 are all positioned on one plane that longitudinally crosses the corresponding shafts. Thus, the input shaft ES and the output shaft DS are positioned on the plane including the first and second lay shafts LSl and LS2.

[106] FIG. 23 illustrates a gear arrangement in which all the three front gears Fl to F3 have the same size, FIG. 24 illustrates a gear arrangement suitable for a vehicle having a small engine in which the first front gear Fl is smaller than the second and third front gears F2 and F3, and FIG. 25 illustrates a gear arrangement suitable for a vehicle having a large engine in which the first front gear Fl is greater than the second and third front gears F2 and F3. In the foregoing embodiments, the output shaft DS is not concentric with the input shaft ES but the output shaft DS and the input shaft ES are spaced apart from each other to be parallel. Accordingly, the axial distance from the output shaft to the first lay shaft LS 1 is different from the axial distance from the output shaft to the second lay shaft LS2. In these cases shown in FIGS. 23 to 25, the ratios of the axial distance between the output shaft DS and the first lay shaft LSI to the axial distance between the output shaft DS and the second lay shaft LS2 are different from one another. Here, the ratios are 3:3.8, 2.8:3.3 and 2.8:4.1, respectively.

[107] Referring to FIGS. 26 to 28, the input shaft ES and the output shaft DS are not positioned on the one plane including the two lay shafts LS 1 and LS2 but are positioned above the plane including the lay shafts LS 1 and LS2. However, the input and output shafts ES and DS are positioned on the same side with respect to the plane including the first and second lay shafts LS 1 and LS2. These cases have different ratios of the axial distance between the output shaft DS and the first lay shaft LSI to the axial distance between the output shaft DS and the second lay shaft LS2. Here, the ratios are 2.8:3.6, 2.7:3.2 and 2.6:3.8, respectively.

[108] Also referring to FIGS. 29 to 31, the input shaft ES and the output shaft DS are positioned on a plane different from the plane including the first and second lay shafts LSI and LS2. In the figures, the input shaft ES is positioned above the plane including the first and second lay shafts LS 1 and LS2, and the output shaft DS is positioned under the plane including the first and second lay shafts LS 1 and LS2. These cases also have different ratios of the axial distance between the output shaft DS and the first lay shaft LS 1 as compared to the axial distance between the output shaft DS and the second lay shaft LS2. Here, the ratios are 2.8:3.6, 2.6:3.2 and 2.6: 3.8, respectively.

[109] As described above, in the present invention, the position of the respective shafts can be altered without limitation. The size of the gearbox and the gear ratio can be freely controlled based on some factors such as the size of an engine and an engine room, a driving mode (e.g., front or rear wheel drive) and an intended use of a vehicle.

[110] FIG. 32 is a schematic plan view for comparing the sizes of gearboxes using conventional transmissions with that of a gearbox using a multi-clutch transmission of the present invention. As shown in the figure, the multi-clutch transmission of the present invention can effectively reduce the volume of a gearbox compared to the conventional manual and automatic transmission and dual-clutch transmissions shown in FIGS. 2 and 3. It is apparent that the size of a gearbox 310 using the transmission of the present invention is much smaller than a gearbox 320 using the conventional dual-clutch transmission shown in FIGS. 2 and 3. Since the volume of the gearbox can be designed to be relatively small using the multi-clutch transmission of the present invention, operational conditions and installation requirements related with arranging a variety of other components in an engine room can be satisfied in a variety of manners. This provides high flexibility in the design and arrangement of the engine room and other parts. In the figure, the reference numeral 305 designates a flywheel, 330 designates the size of the gearbox using a conventional manual transmission, and 340 designates the size of the gearbox using a conventional automatic transmission.

[I l l] While the present invention has been shown and described with reference to certain exemplary embodiments thereof, it will be understood by those skilled in the art that various changes in form and details may be made therein without departing from the spirit and scope of the present invention as defined by the appended claims and their equivalents.

Industrial Applicability The multi-clutch transmission of the present invention has advantageous characteristics such as good power transmission efficiency and high fuel efficiency compared to conventional transmissions. High fuel efficiency can be obtained since the volume of a gearbox can be reduced by controlling the length and width of the gearbox. Accordingly, the present invention has values that can be applied and used in the automobile industry, and it is highly probable that it will be accepted as a next-generation transmission in the art.

Claims

[1] A multi-clutch transmission for transmitting a rotating force from an engine by changing a speed of the rotating force, comprising: an input shaft (210) receiving the rotating force from the engine (200); first and second lay shafts (110, 120) arranged eccentric to an axis of rotation of the input shaft, wherein the first and second lay shafts are spaced apart from each other and parallel; first to third front gears (Fl, F2, F3) for transmitting the rotating force of the input shaft to the first and second lay shafts; first and second clutches (Cl, C2) connected to the first and second lay shafts, respectively, to engage and disengage power; and an output shaft (130) arranged between the first and second lay shafts to be parallel to the first and second lay shafts. [2] The multi-clutch transmission according to claim 1, wherein the first front gear

(Fl) is axially mounted on the input shaft (210), the second front gear (F2) is axially mounted on the first lay shaft (110), and the third front gear (F3) is axially mounted on the second lay shaft (120), wherein the second and third front gears have an equal number of teeth and an equal diameter. [3] The multi-clutch transmission according to claim 1, wherein drive gears of an odd shift stage are arranged on the first lay shaft, and drive gears of an even shift stage are arranged on the second lay shaft. [4] The multi-clutch transmission according to claim 3, further comprising a plurality of output gears arranged on the output shaft (130), wherein the output gears are common gears each of which is meshed with a corresponding one of the drive gears of the odd shift stage and a corresponding one of the drive gears of the even shift stage. [5] The multi-clutch transmission according to claim 4, wherein at least one of the common gears comprises a gear ratio control gear that is provided integratedly with the output gear. [6] The multi-clutch transmission according to claim 3, further comprising a plurality of output gears arranged on the output shaft (130), wherein each of the output gears is meshed with a corresponding one of the drive gears of the odd or even shift stage. [7] The multi-clutch transmission according to claim 1, wherein the second clutch

(C2) is positioned on a plane including the first clutch (Cl). [8] The multi-clutch transmission according to claim 1, wherein the second clutch

(C2) is positioned on a plane different from that including the first clutch (Cl). [9] The multi-clutch transmission according to claim 8, wherein the first and second clutches (Cl, C2) have different diameters. [10] The multi-clutch transmission according to claim 1, further comprising a main clutch (MC) arranged between the engine (200) and the first to third front gears

(Fl, F2, F3) and connected to the input shaft. [11] The multi-clutch transmission according to claim 1, wherein the output shaft

(130) is eccentric to the axis of the input shaft (210). [12] The multi-clutch transmission according to claim 1, further comprising gearbox volume correction gears (CGl, CG2, CG3, CG4) interposed between the first clutch (Cl) and between the second clutch (C2) and the second lay shaft (120). [13] The multi-clutch transmission according to any one of the preceding claims 1 through 12, further comprising a plurality of synchromesh mechanisms arranged on the first lay shaft (110) and the second lay shaft (120), wherein each of the synchromesh mechanisms maintains a pre-selected state coupled with a drive gear of a next shift stage when a selected drive gear of an odd or even shift stage is outputting the rotating force. [14] The multi-clutch transmission according to claim 13, wherein the input shaft

(210) and the output shaft (130) are positioned on a plane including the first and second lay shafts (110, 120). [15] The multi-clutch transmission according to claim 13, wherein the input shaft

(210) and the output shaft (130) are not positioned on a plane including the first and second lay shafts (110, 120).