NL1041642B1 - Automatic transmission with a hydraulic pump. - Google Patents

Abstract

Automatic transmission (1) for transmitting a drive torque between an engine (E) coupled to an input shaft (2) of the transmission (1) and a load (L) coupled to an output shaft of the transmission (1) at a variable transmission ratio, comprising an electro-hydraulic control device provided with an hydraulic pump (10) and an electric motor (41) for driving the pump (10). The transmission (1) further comprises a planetary gearing (50) with a sun gear (51) that is rotationally connected to a drive shaft (55) of the electric motor (41), with one or more planetary gears (52) that are rotatably carried by a planetary carrier (53), which planetary carrier (53) is rotationally connected to the pump (10), and with a ring gear (54) that is rotationally connected to the input shaft (2) via the chain drive 40. In this arrangement of the automatic transmission (1), the pump (10) can favourably be driven by either the engine (E), the electric motor (41) or both.

Description

AUTOMATIC TRANSMISSION WITH A HYDRAULIC PUMP

The present invention relates to an automatic transmission for providing multiple speed ratios between an engine and a load, which transmission is hydraulically actuated and is provided with a hydraulic pump for generating a flow of pressurised hydraulic medium for the hydraulic actuation of such transmission. Typically the pump is driven by the engine via an input shaft of the automatic transmission.

The actuation of the transmission entails at least the switching between two speed ratios of the automatic transmission, i.e. the shifting thereof by the opening and closing of one or more clutches therein. However, in certain kinds of automatic transmission, the said actuation also entails the application of a normal force between two transmission components, so as to enable the transfer of force there between by means of friction. Although several means for generating such normal force are known in the art, it is most commonly generated by pressurising hydraulic medium in the piston-and-cylinder assembly that is associated with at least one of the transmission components. For example in the automatic transmission of the well-known belt-and-pulley continuously variable type, a flexible drive belt is clamped between pulley discs of at least two transmission pulleys. During operation, the pulley discs of each pulley are urged towards each other by applying hydraulic pressure onto at least one of the pulley discs that is thereto slidably mounted on a shaft of the respective pulley. Such a continuously variable transmission is generally known, and e.g. described in the international patent publication WO-A-2006/016797 in the name of Applicant.

As mentioned hereinabove, the transmission pump is typically driven by the engine that also drives the load connected to the transmission. A disadvantage of this particular arrangement is that a maximum flow of hydraulic medium that can be generated by the pump is in principle proportional with the rotational speed of the input shaft, i.e. with the engine speed. Since the pump is required to deliver a sufficient flow of hydraulic medium also at a low engine speed, it normally generates an excess flow at higher engine speeds. This excess pump flow is detrimental to the transmission operation and/or the mechanical efficiency thereof and is thus to be avoided as a matter of principle.

In the art several solutions have been proposed to reduce the said excess pump flow at higher engine speeds, such as the application of a variable flow pump or of a pump system with multiple pumps that can be switched on or off in accordance with the flow requirement. However, such variable pump and/or multiple pumps substantially and disadvantageously increase the complexity and cost of the transmission. To overcome these latter disadvantages, it has been proposed in EP-0764799-A to arrange the pump such that it can be driven either directly by the engine, while generating a flow of hydraulic medium sufficient for the majority of operation conditions, or by an electric motor in case of the operation a higher flow demand. Hereto the driving of the pump by the engine can be overridden by the electric motor, for example by providing a one-way, freewheel or overrunning clutch in the drive line between the engine and the pump.

The present disclosure departs from this latter, known automatic transmission and pump arrangement and sets out to improve the effectiveness thereof. To this end and in accordance with the present disclosure, an epicyclic gearing, i.e. a planetary gearing is included in the pump drive, with the central sun gear and the outer ring gear thereof respectively being drivingly connected to the engine or to the electric motor and with the in-between planet gears being drivingly connected to the pump via the carrier of the planetary gear that rotationally holds the in-between planet gears of the planetary gear.

In the pump arrangement according to the present disclosure, the pump flow is favourably determined by both the engine speed and the speed of the electric motor simultaneously, instead of by either the engine speed or the speed of the electric motor separately. Also in this arrangement, if the engine speed is zero, the pump will still be drivable by the electric motor and vice versa. This thus means that the engine can be switched off, while the transmission can still be actuated by the pump being driven solely by the electric motor and continuing to deliver the required flow of oil. On the other hand, the electric motor can be stopped, if the engine speed is sufficient for the pump to deliver the required flow of oil.

In the latter case, when the electric motor is stopped, it is not enough to merely cut off the electric power, since a rotor of the electric motor will then be rotated in reverse by the engine via the planetary gearing. Therefore, even when stopped, the electric motor is required to generate a torque, which can be done by continuing to electrically power the motor. However, according to the present disclosure, a more efficient and hence more favourable approach is to lock the rotor by mechanical means when the electric motor is stopped. In a preferred embodiment of such mechanical means, these are arranged to prevent rotation in one, i.e. the reverse, direction and to allow rotation in the opposite, i.e. forward direction. For example for this purpose, a worm wheel drive can be provided between the electric motor and the pump or an overrunning clutch can be provided between the rotor and a stator of the electric motor. Since the motor will normally be bolted to a housing of the automatic transmission, such a clutch can be mounted between the said rotor and such housing. In this latter arrangement of the pump drive, the electric motor can be favourably switched-off completely, when the pump flow that is generated by the operation of the engine alone is sufficient for the actuation of the transmission. A further advantage of the pump drive according to the present disclosure is that after the electric motor has been switched-on, it immediately starts to increase the pump flow. Instead, in the known arrangement of EP-0764799-A the electric motor first has to overtake the engine before it is able to increase the pump flow, i.e. the electric motor only drives the pump above a certain rotational speed of its rotor that is determined by the instantaneous engine speed.

The above-described basic features of the present disclosure will now be elucidated by way of example with reference to the accompanying figures.

Figure 1 is a schematic representation of an example of a known hydraulically controlled automatic transmission.

Figure 2 is a symbolic representation of a known and to be improved pump drive.

Figure 3 illustrates a novel pump drive in accordance with the present disclosure.

Figure 4 graphically elucidates the functioning of such novel pump drive.

In the figures, identical references relate to corresponding technical functions or structures, as the case may be.

Figure 1 schematically shows a possible embodiment of an electro-hydraulic control device of an automatic transmission for realising and changing a transmission speed ratio between an input or primary shaft 2 and an output or secondary shaft 3 of the transmission 1. In figure 1 the thick lines indicate hydraulic lines, i.e. passages for hydraulic medium, whereas the dashed lines indicate pressure control lines for the control, i.e. for the biasing of the various hydraulic valves.

The known transmission 1 is intended to be incorporated between an engine E and a load L for varying the transmission speed ratio there between within a continuous range of possible speed ratios. In the shown example, the transmission is a so-called continuously variable transmission 1 that is provided with a drive belt 4 wrapped around and rotationally connecting a primary pulley 5 placed on the input shaft 2 and secondary pulley 6 placed on the secondary shaft 3. The drive belt 4 is frictionally engaged with pulley discs of the respective pulleys 5, 6 by means of respective clamping forces Fp, Fs exerted there between. The primary clamping force Fp is generated by a primary hydraulic pressure Ppri exerted in a pressure chamber 7 of a piston-and-cylinder assembly that is associated with the primary pulley 5, i.e. the primary cylinder 7. The secondary clamping force Fs is generated by a secondary hydraulic pressure Psec exerted in a pressure chamber 8 of a piston-and-cylinder assembly that is associated with the secondary pulley 6, i.e. the secondary cylinder 8.

The known electro-hydraulic control device of the transmission is arranged to realize the primary pressure Ppri and the secondary pressure Psec in a controlled manner. To this end, the control device comprises a hydraulic pump 10 for generating a flow of hydraulic medium from a reservoir at low pressure to a main hydraulic line 12 at high(-er) pressure. The pressure of the hydraulic medium in this main line 12, i.e. the pump or line pressure Pline, is controlled by means of a pressure control valve, i.e. a line pressure valve 13. This line pressure valve 13 is provided with valve biasing means including a spring 13a, a valve actuator 13c and a pressure-feedback line 13b that together control the line pressure Pline. From this line pressure Pline, the primary pressure Ppri and the secondary pressure Psec are derived by means of a primary pressure valve 20 and a secondary pressure valve 30 respectively. The primary pressure valve 20 is interposed between the main line 12 and a primary hydraulic branch 21 connected to the primary cylinder 7 and the secondary pressure valve 30 is interposed between the main line 12 and a secondary hydraulic branch 31 connected to the secondary cylinder 8. Both the primary pressure valve 20 and the secondary pressure valve 30 are provided with valve biasing means of their own, respectively including a spring 20b; 30b, a valve actuator 20c; 30c and a pressure-feedback line 20a; 30a, which respective valve biasing means function similar to the said valve biasing means of the line pressure valve 13.

Further hydraulic functions AF of the transmission are supplied with hydraulic medium from an auxiliary line 14 of the known control device that receives such hydraulic medium from the line pressure valve 13. The pressure of the hydraulic fluid in the auxiliary line 14, i.e. the auxiliary pressure Paux, is controlled by means of a further pressure control valve, i.e. an auxiliary pressure valve 15 with own valve biasing means 15a, 15b.

Finally, the known control device also supplies hydraulic medium to the moving parts MP of the transmission 1, such as the drive belt 4 and shaft bearings, for the lubrication and the cooling thereof. In the control device of figure 1, yet a further hydraulic line 16 and pressure control valve, i.e. a lubrication pressure valve 17 are included for this purpose.

It is noted that many alternative arrangements of the automatic transmission and of the electro-hydraulic control device thereof are known. The presently illustrated continuously variable transmission 1 and electro-hydraulic control device are a mere example thereof.

In figure 2 a known arrangement for driving the pump 10 in the known transmission is illustrated. In such known pump drive the pump 10 is driven by the engine E via a drive chain 40 mounted on sprockets that drivingly connects the pump 10 to the input shaft 2 of the transmission 1. Furthermore, an electric motor 41 is included in the pump drive, which motor 41 is arranged to drive the pump 10 instead of the engine E. In this known dual pump drive an overrunning clutch 42 is included between the input shaft 2 and the pump 10, which freewheel clutch 42 is arranged to allow the electric motor 41 to drive the pump 10 at increased rotational speed relative to the rotational speed determined by the engine E. By such (relatively) increased rotational speed, the pump flow from the reservoir 11 to the above-described hydraulic components 12-31 of the control device -and ultimately the primary cylinder 7 and the secondary cylinder 8- is increased accordingly.

At least one drawback of the known pump drive of figure 2 is that the electric motor 41 has to be of considerable size and power in order to solely drive the pump 10 to accommodate the maximum flow and pressure demand of the transmission 1. The present disclosure aims to mitigate such drawback. Hereto, a novel pump drive is provided that is illustrated in figure 3 and wherein the engine E and the electric motor 41 work together to jointly accommodate the said maximum flow and pressure demand.

The central component of the novel pump drive is an epicylic or planetary gearing 50 that is as such well-known. The planetary gearing 50 comprises a central or sun gear 51, one or more planetary gears 52 that are rotatably carried by a planetary carrier 53 and an annulus or ring gear 54. The sun gear 51, the planetary carrier 53 and the ring gear 54 are arranged concentrically rotatable, with the planetary gears 52 meshing with the sun gear 51 and with the ring gear 54. The ring gear 54 is rotationally connected to and can thus be driven by the input shaft 2 via the chain drive 40. The sun gear 51 is rotationally connected to and can thus be driven by a drive shaft 55 of the electric motor 41. The planetary carrier 53 is rotationally connected to and can thus drive the pump 10.

Furthermore, the novel pump drive comprises an overrunning clutch 56 that is symbolically indicated in figure 3 by a diode symbol 56 and that is arranged between the drive shaft 55 of the electric motor 41 and a non-rotating part 57 such as a housing 57 of the transmission 1. The overrunning clutch 56 allows the drive shaft 55 of the electric motor 41 to rotate and drive the pump 10, but prevents a reverse, i.e. counter rotation thereof, i.e. provides a counter torque to stop the sun gear 51 from rotating in reverse when the electric motor 41 is switched off. The practical operation of the novel pump drive of figure 3 is elucidated with reference to figure 4.

Figure 4 is a diagram wherein the rotational speeds ω of the components of the planetary gearing 50 are indicated on the three horizontal X-axes. The rotational speed ω54 of the ring gear 54 that is determined by and proportional to the rotational speed of the engine E is plotted on the uppermost X-axis. The rotational speed ω53 of the planet carrier 53 that determines the rotational speed of the pump 10 is plotted on the middle X-axis. The rotational speed ωει of the sun gear 51 that is determined by and proportional to the rotational speed of the electric motor 41 is plotted on the lowermost X-axis in figure 4. The vertical separation between these three X-axes reflects a speed ratio A between the ring gear 54 and the planetary carrier 53 and between the planetary carrier 53 and the sun gear 51, i.e. speed ratio B, respectively.

The dashed lines D1, D2 in figure 4 represent a situation wherein the pump 10 is driven solely by the engine E, i.e. wherein the electric motor 41 is stopped. Obviously, in this case the rotational speed u)si of the electric motor 41 and thus of the sun gear 51 is zero and the rotational speed ω53 of the planetary carrier 53 driving the pump 10 is proportional with the rotational speed u>54 of the ring gear 54 driven by the engine E. It may be noted that such proportionality is determined by the said speed ratios A and B of the planetary gearing 50 as a whole. Thus, as the speed of the ring gear 54 increases (e g. from 1-1054 to 2-0)54), the speed of the planetary gear 53 driving the pump 10 increases proportionally (e.g. from 1-ius3 to 2-u)53). Furthermore, it may be noted that in this situation, the electric motor 41, in particular the drive shaft 55 thereof, is to be blocked from rotating in reverse for example by providing a mechanical connection to a nonrotating part 57 of the transmission 1.

When necessary, i.e. when required by and/or for the hydraulic actuation of the transmission, the electric motor 41 is powered to rotate the sun gear 51, whereby the rotational speed ω53 of the planetary carrier 53 that drives the pump 10 is increased. The dashed-dotted line D3 in figure 4 represents such situation, wherein the rotational speed ω54 of the ring gear 54 driven by the engine E is relatively low (e.g. 1-(054) and the rotational speed of the planetary gear 53 driving the pump 10 is boosted (e.g. from I-0053 to 3-0053) by the activation of the electric motor 41 (e.g. at 1-0054).

Furthermore, the novel pump drive also allows the engine E to be shut-off completely, as for example may be desired to minimise fuel consumption. This situation is represented by the dotted line D4 in figure 4. Obviously, in this case the rotational speed CO54 of the engine E and thus of the ring gear 54 is zero and the rotational speed 0053 of the planetary carrier 53 driving the pump 10 (e.g. 4-0053) is proportional with the rotational speed (Ο51 of the sun gear 51 driven by the electric motor 41 (e.g. 2-0051).

The present disclosure, in addition to the entirety of the preceding description and all details of the accompanying figures, also concerns and includes all the features of the appended set of claims. Bracketed references in the claims do not limit the scope thereof, but are merely provided as non-binding examples of the respective features. The claimed features can be applied separately in a given product or a given process, as the case may be, but it is also possible to apply any combination of two or more of such features therein.

The invention(s) represented by the present disclosure is (are) not limited to the embodiments and/or the examples that are explicitly mentioned herein, but also encompasses amendments, modifications and practical applications thereof, in particular those that lie within reach of the person skilled in the relevant art.

Claims (5)

An automatic transmission (1) for transmitting, with a variable transmission ratio, a driving power between a motor (E) coupled or connectable to an input shaft (2) of the transmission (1) and an output shaft (E) (3) load (L) coupled or connectable from the transmission (1), in particular in a drive line of a vehicle, provided with at least one pressure chamber (7; 8), in which a hydraulic pressure is adjusted by means of a hydraulic control device (Ppri; Psec), of a hydraulic pump (10) and of an electric motor (41), which pump (10) is drivable by the motor (E) and / or by the electric motor (41), characterized in that the transmission (1) is also provided with a planetary gear set (50) with a sun gear (51), with a ring gear (54) and with a number of planet wheels (52) mounted rotatably in a planet carrier (53), of which either the sun gear (51) or the ring gear (54) on the input shaft (2) is g is coupled and drivable by the motor (E), whose sun gear (51) or ring wheel (54) not coupled to the input shaft (2) is coupled to a drive shaft (55) of the electric motor (41) and is drivable by said motor electric motor (41) and of which the planet carrier (53) is coupled to the pump (10) and can drive it. Automatic transmission (1) according to claim 1, characterized in that means (56) are provided therein for preventing rotation of the drive shaft (55) of the electric motor (41) in a pump drive (10) opposite direction of rotation. Automatic transmission (1) according to claim 2, characterized in that said means (56) comprise a freewheel clutch (56) which is between the drive shaft (55) and non-rotating part of the transmission (1), such as a housing of the transmission (1) or of the electric motor (41) itself. Automatic transmission (1) according to claim 1, 2 or 3, characterized in that it is provided with two, with the aid of, respectively, a primary hydraulic pressure (Ppri) in a primary pressure chamber (7) and a secondary hydraulic pressure ( Psec) in a secondary pressure chamber (8), powered pulleys (5, 6), around which a drive belt (4) is laid and which are respectively arranged on the input shaft (2) and the output shaft (3) of the transmission (1) . A method for driving a hydraulic pump (10) of an automatic transmission (1) for passing, with a variable transmission ratio, a driving power between a coupled to an input shaft (2) of the transmission (1) or connectable motor (E) and a load (L) coupled or connectable to an output shaft (3) of the transmission (1) and provided with at least one pressure chamber (7; 8), in which a hydraulic pressure is applied by means of a hydraulic control device adjusted (Ppri; Psec), of an electric motor (41) and of a planetary gear set (50) with a sun gear (51), with a ring wheel (54) and with a number of planetary wheels (52) rotatably mounted in a planet carrier (53) ) therebetween, of which either the sun gear (51) or the ring gear (54) is coupled to the input shaft (2), the sun gear (51) or ring gear (54) of which is not coupled to the input shaft (2) to a drive shaft ( 55) of the electric motor (41) and where of the planet carrier (53) is coupled to the pump (10), wherein the pump (10) is optionally driven by the motor (E), by the electric motor (41) or by both the motor (E) and the electric motor ( 41).