JP4442417B2 - Vehicle lubrication control device - Google Patents

Description

  The present invention provides a power transmission having a first motor that functions as a generator, a second motor that functions as a driving force source for traveling, and a starting clutch that connects the first motor and the second motor so that they can be connected and disconnected. More particularly, the present invention relates to a technique for controlling the circulating supply of lubricating fluid to the power transmission device.

  A vehicle configured to circulate and supply a lubricating fluid to a power transmission device having an electric motor and to cool the electric motor with the lubricating fluid is well known. In such a vehicle, in order to improve the cooling performance of the electric motor, it is only necessary to increase the lubrication pressure and increase the flow rate of the lubricating fluid. However, if the lubrication pressure is increased uniformly and the flow rate is constantly increased, the fuel efficiency is improved. Causes it to get worse.

  Therefore, in the vehicle described in Patent Document 1, a power transmission device having a first motor coupled to the engine and a second motor coupled to the engine and the first motor via a start clutch, and to the power transmission device A fluid circulation supply device that circulates and supplies the lubricating fluid, and increases the flow rate of the lubricating fluid when the temperature of the first electric motor or the second electric motor is equal to or higher than a predetermined temperature.

JP 2002-147484 A JP 2002-340167 A JP 2004-100580 A

  By the way, in the vehicle as described in the above-mentioned Patent Document 1, it is directly driven by the electric energy generated by the first electric motor driven by the power of the engine or indirectly by the electric energy once charged in the power storage device. In addition, the second electric motor is rotationally driven to generate an output. For example, as one method of starting the vehicle, there is a method of starting the vehicle exclusively using the output of the second electric motor as a driving force in a state where the starting clutch is released. On the other hand, when the motor malfunctions or when the motor performance drops due to low temperatures, as another starting method, the starting clutch is slip-engaged and the slip amount is controlled so that the engine output is gradually transmitted to the driving wheels. There is a so-called friction start method for starting a vehicle.

  However, if the slip engagement time of the starting clutch is prolonged, the slip amount is large, or the temperature of the lubricating fluid is high at the time of the friction start, the durability of the starting clutch may be lowered. In particular, when such a friction start is executed, the first and second electric motors are not in an operating state, so that the temperature rise is suppressed as compared with the operating state. If the flow rate of the lubricating fluid is controlled in accordance with the temperatures of the first and second electric motors, the flow rate may not be properly controlled.

  Alternatively, it is conceivable that the starting clutch released in the case of a starting method other than the friction start is put into a slip engagement state, and the vehicle is started using a part of the output of the engine as a driving force in addition to the output of the second electric motor. In such a case, the output of the second motor is reduced by the amount of the driving force due to the output of the engine as compared with the case where the necessary driving force is obtained exclusively by the output of the second motor. The amount of heat generation is reduced and the temperature rise is suppressed, or the electric energy required for driving the second motor is reduced, and the amount of power generated by the first motor is reduced and the temperature rise is suppressed. Accordingly, even in such a case, if the flow rate of the lubricating fluid is controlled according to the temperature of the electric motor as in Patent Document 1, the flow rate may not be appropriately controlled.

  The present invention has been made against the background of the above circumstances. The object of the present invention is to provide a power transmission device having a starting clutch that connects and disconnects the first motor and the second motor, and its power. An object of the present invention is to provide a vehicle lubrication control device that appropriately cools a first electric motor and / or a second electric motor with a lubricating fluid in a vehicle including a fluid circulation supply device that circulates and supplies a lubricating fluid to a transmission device.

The gist of the first invention for achieving the object is as follows: (a) a first motor that functions as a generator by being driven to rotate by an engine, and a motor that is driven to rotate by the generated energy of the first motor; A power transmission device having a second electric motor that functions as a driving power source for traveling, a starting clutch that connects the engine, the first electric motor, and the second electric motor so as to be connectable / disengageable, and for lubricating the power transmission device A fluid circulation supply device for circulating and supplying fluid and cooling the first electric motor and the second electric motor with the lubricating fluid, wherein (b) the engine and the second electric motor Based on the torque capacity of the starting clutch during operation, the fluid circulation supply device controls the lubrication pressure of the lubricating fluid circulated and supplied to the power transmission device. The pressure control means, seen including, (c) the lubricating pressure control means, degree of driving force based on the output of the engine to be transmitted to the drive wheels in accordance with the torque capacity of the starting clutch is relatively large, The lubricating pressure of the lubricating fluid is corrected to be small .

If it does in this way, the power transmission device which has the starting clutch which connects the 1st electric motor and the 2nd electric motor so that connection and disconnection is possible, circulates and supplies the fluid for lubrication to the power transmission device, and the 1st by the fluid for lubrication In a vehicle including a motor and a fluid circulation supply device that cools the second motor, the fluid circulation supply device circulates to the power transmission device based on the torque capacity of the starting clutch when the engine and the second motor are in operation. A lubricating pressure control means for controlling a lubricating pressure of the supplied lubricating fluid, and the driving force based on the output of the engine transmitted to the driving wheel according to the torque capacity of the starting clutch is relatively increased. , the lubricating pressure of the lubricating fluid by the lubricating pressure control means are less corrected, out of the engine according to the torque capacity of the starting clutch Calorific output is relatively unchanged required for driving force component to obtain the same driving force by relative change based on the (temperature) of the first and second electric motors which changes relatively The flow rate required for cooling can be obtained appropriately. Thereby, durability of a 1st electric motor and a 2nd electric motor improves.

  Here, preferably, in the invention according to claim 2, the lubricating pressure is determined based on a temperature state of the first electric motor or the second electric motor. In this way, the flow rate required for cooling the first motor and / or the second motor can be obtained more appropriately, and the durability of the first motor and / or the second motor can be further improved.

  For example, the higher the temperature that is one of the temperature states of the first motor or the second motor, the higher the necessity of cooling the first motor or the second motor itself. If the flow rate of the fluid is increased, the cooling effect is further enhanced. Alternatively, if the temperature change that is one of the temperature states of the first motor or the second motor is larger, the temperature of the first motor or the second motor is not so high that the lubricating pressure must be increased. However, since the possibility of the temperature becoming relatively high and the necessity of cooling the first motor or the second motor itself is increased, the lubricating pressure is increased to increase the flow rate of the lubricating fluid. As a result, the cooling effect is further enhanced.

  Preferably, in the invention according to claim 3, the lubricating pressure is determined based on a heat generation state of the first electric motor or the second electric motor. In this way, the flow rate required for cooling the first motor and / or the second motor can be obtained more appropriately, and the durability of the first motor and / or the second motor can be further improved.

  For example, the higher the heat generation amount that is one of the heat generation states of the first motor or the second motor, or the larger the change in the heat generation amount that is one of the heat generation states, the first Even when the temperature of the electric motor or the second electric motor is not so high that the lubrication pressure has to be increased, the possibility that the temperature is increased is relatively high and the first electric motor or the second electric motor itself is cooled. Since the necessity increases, the cooling effect can be further enhanced by increasing the lubricating pressure and increasing the flow rate of the lubricating fluid.

  Preferably, in the invention according to claim 4, the lubricating pressure is determined based on a temperature state of the lubricating fluid. In this way, the flow rate required for cooling the first motor and / or the second motor can be obtained more appropriately, and the durability of the first motor and / or the second motor can be further improved.

  For example, the higher the temperature that is one of the temperature states of the lubricating fluid, the higher the temperature of the first motor or the second motor itself, or the higher the temperature of the first motor or the second motor. Therefore, if the lubricating pressure is increased to increase the flow rate of the lubricating fluid, the cooling effect can be further enhanced. Alternatively, even if the temperature of the lubricating fluid is not so high that the temperature of the lubricating fluid must be increased as the temperature change, which is one of the temperature states of the lubricating fluid, is larger. Since the possibility of increasing becomes relatively high, the cooling effect can be further enhanced by increasing the lubricating pressure and increasing the flow rate of the lubricating fluid.

The subject matter of the fifth invention for achieving the above object is as follows: (a) a first motor that functions as a generator by being rotationally driven by the engine, and rotationally driven by the generated energy of the first motor; And a power transmission device having a second motor that functions as a driving power source for traveling, a starting clutch that connects the first motor and the second motor so as to be connectable and disconnectable, and the power transmission device. A fluid circulation supply device for circulating and supplying a lubricating fluid and cooling the first electric motor and the second electric motor with the lubricating fluid, wherein (b) the engine and the second controlling the lubrication flow rate of the lubricating fluid is circulated and supplied to the power transmission device by the fluid circulation and supply system based on the torque capacity of the starting clutch during the operation of the electric motor Lubrication flow rate control means that, seen including, (c) the lubricant flow rate control means, the driving force based on the output of the engine to be transmitted to the drive wheels in accordance with the torque capacity of the starting clutch is relatively large The purpose is to correct the lubricating flow rate of the lubricating fluid smaller .

If it does in this way, the power transmission device which has the starting clutch which connects the 1st electric motor and the 2nd electric motor so that connection and disconnection is possible, circulates and supplies the fluid for lubrication to the power transmission device, and the 1st by the fluid for lubrication In a vehicle including a motor and a fluid circulation supply device that cools the second motor, the fluid circulation supply device circulates to the power transmission device based on the torque capacity of the starting clutch when the engine and the second motor are in operation. A lubricating flow rate control means for controlling a lubricating flow rate of the supplied lubricating fluid, and the driving force based on the output of the engine transmitted to the driving wheels according to the torque capacity of the starting clutch is relatively increased. the the lubricating flow of lubricating fluid is reduced corrected by the lubricant flow rate control means, ene according to the torque capacity of the starting clutch Calorific output is relatively unchanged required for driving force component based on the output of the down obtain a same driving force that changes relatively first electric motor (temperature) is relatively changed and the 2 The flow rate required for cooling the motor can be obtained appropriately. Thereby, durability of a 1st electric motor and a 2nd electric motor improves.

  Preferably, in the invention according to claim 6, the lubricating flow rate is determined based on a temperature state of the first electric motor or the second electric motor. In this way, the flow rate required for cooling the first motor and / or the second motor can be obtained more appropriately, and the durability of the first motor and / or the second motor can be further improved.

  For example, the higher the temperature that is one of the temperature states of the first motor or the second motor, the higher the need to cool the first motor or the second motor itself. If the flow rate of the fluid is increased, the cooling effect is further enhanced. Alternatively, if the temperature change that is one of the temperature states of the first motor or the second motor is larger, the temperature of the first motor or the second motor is not so high that the lubrication flow rate must be increased. However, since the possibility of the temperature becoming relatively high and the necessity of cooling the first motor or the second motor itself is increased, the lubricating flow rate is increased to increase the flow rate of the lubricating fluid. As a result, the cooling effect is further enhanced.

  Preferably, in the invention according to claim 7, the lubricating flow rate is determined based on a heat generation state of the first electric motor or the second electric motor. In this way, the flow rate required for cooling the first motor and / or the second motor can be obtained more appropriately, and the durability of the first motor and / or the second motor can be further improved.

  For example, the higher the heat generation amount that is one of the heat generation states of the first motor or the second motor, or the larger the change in the heat generation amount that is one of the heat generation states, the first Even when the temperature of the electric motor or the second electric motor is not so high that the lubrication flow rate has to be increased, the possibility that the temperature becomes higher is relatively high, and the first electric motor or the second electric motor itself is cooled. Since the necessity increases, the cooling effect can be further enhanced by increasing the lubricating flow rate and increasing the flow rate of the lubricating fluid.

  Preferably, in the invention according to claim 8, the lubricating flow rate is determined based on a temperature state of the lubricating fluid. In this way, the flow rate required for cooling the first motor and / or the second motor can be obtained more appropriately, and the durability of the first motor and / or the second motor can be further improved.

  For example, the higher the temperature that is one of the temperature states of the lubricating fluid, the higher the temperature of the first motor or the second motor itself, or the higher the temperature of the first motor or the second motor. Therefore, if the lubricating flow rate is increased to increase the lubricating fluid flow rate, the cooling effect can be further enhanced. Alternatively, even if the temperature of the lubricating fluid is not so high that the temperature of the lubricating fluid must be increased as the temperature change, which is one of the temperature states of the lubricating fluid, is larger. Since the possibility of increase becomes relatively high, the cooling effect can be further enhanced by increasing the lubrication flow rate and increasing the flow rate of the lubricating fluid.

  Here, preferably, the power transmission device is provided with a transmission. As the transmission, a planetary gear type multi-stage transmission in which a gear stage is switched by selectively connecting rotating elements of a plurality of planetary gear units by a hydraulic friction engagement device, a transmission functioning as a power transmission member A belt-type continuously variable transmission in which a belt is wound around a pair of variable pulleys whose effective diameters are variable and the gear ratio is continuously changed steplessly, a pair of cones and their shafts rotated around a common axis Traction whose rotation ratio is variable by the fact that a plurality of rotatable rollers that intersect the center of rotation rotate between the pair of cones and change the angle of intersection between the center of rotation of the rollers and the shaft center. Type continuously variable transmission, a plurality of pairs of transmission gears that are always meshed between two shafts, and one of these pairs of transmission gears is selected by a synchronization device driven by a hydraulic actuator or the like Or the like can be used synchromesh type parallel two-shaft transmission that the power transmission state.

  The mounting posture of the transmission with respect to the vehicle may be an FR (front engine / front drive) FF vehicle in which the transmission axis is in the width direction of the vehicle. Front engine / rear drive) Vertical type such as a vehicle may be used.

  Further, the planetary gear type multi-stage transmission may be any gear that can alternatively achieve a plurality of gear stages, for example, forward 4 stages, forward 5 stages, forward 6 stages, forward 7 stages, forward 8 stages. Various multi-stage automatic transmissions such as can be used.

  Preferably, the starting clutch is a multi-plate or single-plate hydraulic friction engagement device that is engaged by a hydraulic actuator (hydraulic cylinder). An oil pump that supplies hydraulic oil for engaging the starting clutch and the hydraulic friction engagement device in the planetary gear type multi-stage transmission is directly driven by engine power such as a mechanical oil pump. The hydraulic oil may be discharged, but it may be driven by a dedicated electric motor disposed separately from the engine, such as an electric hydraulic pump.

  The hydraulic oil is also a lubricating fluid that is circulated and supplied to the power transmission device by the fluid circulation supply device.

  Hereinafter, embodiments of the present invention will be described in detail with reference to the drawings.

  FIG. 1 is a skeleton diagram illustrating the configuration of a vehicle drive device (hereinafter referred to as drive device) 6 as a power transmission device provided in a vehicle to which the present invention is applied. This drive device 6 is mechanically connected to a first motor generator MG1 as a first motor, an engine 8 and an automatic transmission 10 on a common axis in a transmission case 12 as a non-rotating member attached to a vehicle body. A starting clutch Ci that can be connected and disconnected in such a manner, a second motor generator MG2 as a second electric motor, and a stepped automatic transmission 10 are sequentially arranged. The automatic transmission 10 mainly includes an input shaft 16 and a first planetary gear unit 18 which are mechanically connected to a crankshaft 9 of an engine 8 which is an internal combustion engine such as a gasoline engine or a diesel engine exclusively via a starting clutch Ci. The first transmission unit 20, the second planetary gear unit 22, the second planetary gear unit 22 and the third planetary gear unit 24, and the output shaft 28 are sequentially arranged. The rotation of 16 is changed and output from the output shaft 28. The input shaft 16 functions as an output rotation member of the starting clutch Ci, and also functions as an input rotation member of the automatic transmission 10. Further, the output shaft 28 corresponds to an output rotating member of the automatic transmission 10, and for example, as shown in FIG. 7, the left and right sides are sequentially passed through a differential gear device (final reduction gear) 30, a pair of axles 31, and the like. The drive wheel 32 is rotationally driven. The first motor generator MG1 is directly operatively connected to the engine 8 (crankshaft 9), and the second motor generator MG2 is directly operatively connected to the input shaft 16.

  As described above, in the driving device 6 of this embodiment, the crankshaft 9 and the input shaft 16 can be mechanically connected, that is, directly connected (hereinafter referred to as direct connection) via the starting clutch Ci. This direct connection means that the connection is made without using a hydraulic power transmission device such as a torque converter or a fluid coupling. For example, the connection through a pulsation absorbing damper (vibration damping device) is included in this direct connection. Since the drive device 6 is configured symmetrically with respect to its axis, the lower side is omitted in the skeleton diagram of FIG.

  The first planetary gear unit 18 is a double-pinion type planetary gear unit, and includes a sun gear S1, a plurality of pairs of pinion gears P1 that mesh with each other, a carrier CA1 that supports the pinion gears P1 so as to rotate and revolve, and a sun gear S1 via the pinion gears P1. A ring gear R1 meshing with the ring gear R1. The carrier CA1 is coupled to the input shaft 16 and driven to rotate, and the sun gear S1 is fixed to the transmission case 12 so as not to rotate. The ring gear R <b> 1 functions as an intermediate output member, is rotated at a reduced speed with respect to the input shaft 16, and transmits the rotation to the second transmission unit 26. In the present embodiment, the path for transmitting the rotation of the input shaft 16 to the second transmission unit 26 at the same speed is the first intermediate output path for transmitting the rotation at a predetermined constant speed ratio (= 1.0). The first intermediate output path PA1 includes a direct connection path PA1a that transmits rotation from the input shaft 16 to the second transmission unit 26 without passing through the first planetary gear unit 18, and a first planetary gear from the input shaft 16. There is an indirect path PA1b that transmits the rotation to the second transmission unit 26 via the carrier CA1 of the device 18. Further, the transmission ratio from the input shaft 16 to the second transmission 26 via the carrier CA1, the pinion gear P1 disposed on the carrier CA1, and the ring gear R1 is larger than the first intermediate output path PA1 (> 1). .0) is a second intermediate output path PA2 that transmits the rotation of the input shaft 16 at a reduced speed (deceleration).

  The second planetary gear device 22 is a single pinion type planetary gear device, and includes a sun gear S2, a pinion gear P2, a carrier CA2 that supports the pinion gear P2 so as to be capable of rotating and revolving, and a ring gear R2 that meshes with the sun gear S2 via the pinion gear P2. I have. The third planetary gear unit 24 is a double pinion type planetary gear unit, which includes a sun gear S3, a plurality of pairs of pinion gears P2 and P3 that mesh with each other, a carrier CA3 that supports the pinion gears P2 and P3 so as to be capable of rotating and revolving, a pinion gear P2 and A ring gear R3 meshing with the sun gear S3 via P3 is provided.

  In the second planetary gear unit 22 and the third planetary gear unit 24, the carriers CA2 and CA3 that rotatably support the pinion gear P2 and the ring gears R2 and R3 are shared with each other, so that four rotating elements RM1 to RM4 are configured. ing. That is, the first rotating element RM1 is configured by the sun gear S2 of the second planetary gear unit 22, and the carrier CA2 of the second planetary gear unit 22 and the carrier CA3 of the third planetary gear unit 22 are integrally connected to each other to perform the second rotation. The element RM2 is configured, and the ring gear R2 of the second planetary gear unit 22 and the ring gear R3 of the third planetary gear unit 24 are integrally connected to each other to configure the third rotating element RM3, and the sun gear of the third planetary gear unit 24 is configured. The fourth rotation element RM4 is configured by S3.

  The first rotating element RM1 (sun gear S2) is selectively connected to the transmission case 12 via the first brake B1 and stopped, and the first planetary gear unit 18 which is an intermediate output member via the third clutch C3. Ring gear R1 (that is, the second intermediate output path PA2) is selectively connected to the carrier CA1 of the first planetary gear unit 18 (that is, the indirect path PA1b of the first intermediate output path PA1) via the fourth clutch C4. Is selectively linked. The second rotation element RM2 (carriers CA2 and CA3) is selectively connected to the transmission case 12 via the second brake B2 and stopped, and the input shaft 16 (that is, the first shaft 16 via the second clutch C2). It is selectively connected to the direct connection path PA1a) of the intermediate output path PA1. The third rotation element RM3 (ring gears R2 and R3) is integrally connected to the output shaft 28 to output rotation. The fourth rotation element RM4 (sun gear S3) is connected to the ring gear R1 via the first clutch C1. The starting clutch Ci, the brakes B1 and B2, and the clutches C1 to C4 are all hydraulic friction engagement devices such as a multi-plate type that are frictionally engaged by a hydraulic actuator (hydraulic cylinder).

  FIG. 2 is a collinear diagram in which the rotational speeds of the rotary elements of the first transmission unit 20 and the second transmission unit 26 can be represented by straight lines. The lower horizontal line indicates the rotational speed “0”. The horizontal line indicates the rotational speed “1.0”, that is, the same rotational speed as the input shaft 16. Further, each vertical line of the first transmission unit 20 represents the sun gear S1, the ring gear R1, and the carrier CA1 in order from the left side, and these intervals are the gear ratio ρ1 (= sun gear S1 of the first planetary gear unit 18). The number of teeth / the number of teeth of the ring gear R1). FIG. 2 shows a case where the gear ratio ρ1 = 0.463, for example. The four vertical lines of the second transmission unit 26 indicate, in order from the left side, the first rotating element RM1 (sun gear S2), the second rotating element RM2 (carrier CA2 and carrier CA3), and the third rotating element RM3 (ring gear R2 and ring gear). R3), the fourth rotation element RM4 (sun gear S3), and their intervals are determined according to the gear ratio ρ2 of the second planetary gear unit 22 and the gear ratio ρ3 of the third planetary gear unit 24. FIG. 2 shows a case where the gear ratio ρ2 = 0.463 and ρ3 = 0.415, for example.

  As is clear from this nomograph, the first clutch C1 and the second brake B2 are engaged, and the fourth rotating element RM4 rotates at a reduced speed with respect to the input shaft 16 via the first transmission 20. When the rotation of the second rotation element RM2 is stopped, the third rotation element RM3 connected to the output shaft 28 is rotated at the rotation speed indicated by “1st”, and the largest transmission ratio (= input shaft 16). ) / (Rotational speed of the output shaft 28)) is established.

  The first clutch C1 and the first brake B1 are engaged, and the fourth rotating element RM4 is decelerated and rotated with respect to the input shaft 16 via the first transmission unit 20, and the first rotating element RM1 stops rotating. Then, the third rotation element RM3 is rotated at the rotation speed indicated by “2nd”, and the second speed “2nd” having a smaller gear ratio than the first speed “1st” is established.

  The first clutch C1 and the third clutch C3 are engaged, and the fourth rotation element RM4 and the first rotation element RM1 are decelerated and rotated with respect to the input shaft 16 via the first transmission unit 20 to perform the second shift. When the part 26 is rotated integrally, the third rotation element RM3 is rotated at the rotation speed indicated by “3rd”, and the third shift stage “3rd” having a smaller speed ratio than the second shift stage “2nd” is established. It is done.

  The first clutch C1 and the fourth clutch C4 are engaged, the fourth rotating element RM4 is decelerated and rotated with respect to the input shaft 16 via the first transmission unit 20, and the first rotating element RM1 is input to the input shaft. When it is rotated integrally with 16, the third rotation element RM3 is rotated at the rotational speed indicated by “4th”, and the fourth shift stage “4th” having a smaller speed ratio than the third shift stage “3rd” is established. .

  When the first clutch C1 and the second clutch C2 are engaged, the fourth rotation element RM4 is rotated at a reduced speed with respect to the input shaft 16 via the first transmission unit 20, and the second rotation element RM2 is input to the input shaft 16. The third rotation element RM3 is rotated at the rotation speed indicated by “5th”, and the fifth shift stage “5th” having a smaller gear ratio than the fourth shift stage “4th” is established.

  When the second clutch C2 and the fourth clutch C4 are engaged and the second transmission unit 26 is rotated integrally with the input shaft 16, the third rotational element RM3 is rotated at the rotational speed indicated by "6th", that is, with the input shaft 16. The sixth shift stage “6th”, which is rotated at the same rotational speed and has a smaller speed ratio than the fifth shift stage “5th”, is established. The gear ratio of the sixth gear stage “6th” is 1.

  The second clutch C2 and the third clutch C3 are engaged, and the first rotating element RM1 is decelerated and rotated with respect to the input shaft 16 via the first transmission unit 20, and the second rotating element RM2 is input to the input shaft. When it is rotated integrally with 16, the third rotation element RM3 is rotated at the rotational speed indicated by “7th”, and the seventh shift stage “7th” having a smaller gear ratio than the sixth shift stage “6th” is established. .

  When the second clutch C2 and the first brake B1 are engaged, the second rotating element RM2 is rotated integrally with the input shaft 16, and when the first rotating element RM1 is stopped, the third rotating element RM3 is The eighth speed “8th” is established at a rotational speed indicated by “8th” and has a smaller gear ratio than the seventh speed “7th”.

  When the third clutch C3 and the second brake B2 are engaged, the first rotating element RM1 is rotated at a reduced speed via the first transmission unit 20, and the second rotating element RM2 is stopped from rotating. The third rotation element RM3 is reversely rotated at the rotation speed indicated by “Rev1”, and the first reverse shift stage “Rev1” having the largest speed ratio in the reverse rotation direction is established. When the fourth clutch C4 and the second brake B2 are engaged, the first rotation element RM1 is rotated integrally with the input shaft 16, the second rotation element RM2 is stopped, and the third rotation element RM3 is “ The second reverse shift speed “Rev2”, which is reversely rotated at the rotation speed indicated by “Rev2” and has a smaller gear ratio than the first reverse shift speed “Rev1”, is established. The first reverse speed “Rev1” and the second reverse speed “Rev2” correspond to the first speed and the second speed in the reverse rotation direction, respectively.

  FIG. 3 is an operation table for explaining the engagement elements and the gear ratios when the above gear positions are established. “◯” indicates the engaged state, and the blank is released. The gear ratio of each gear stage is appropriately determined by the gear ratios ρ1 to ρ3 of the first planetary gear device 18, the second planetary gear device 22, and the third planetary gear device 24, for example, ρ1 = 0.463, ρ2 = 0. .463, ρ3 = 0.415, the value of the gear ratio step (the gear ratio between the gears) is substantially appropriate, and the total gear ratio width (= 4.495 / 0.683) is also obtained. The gear ratio of the reverse gears “Rev1” and “Rev2” is also appropriate, and an appropriate gear ratio characteristic is obtained as a whole.

  As described above, the automatic transmission 10 according to the present embodiment includes the first transmission unit 20 having the two intermediate output paths PA1 and PA2 having different transmission ratios and the second transmission unit 26 having the two planetary gear units 22 and 24. Since the forward shift 8-speed gear stage is achieved by switching the engagement of the four clutches C1 to C4 and the two brakes B1 and B2, the structure is reduced in size and the mountability to the vehicle is improved. As can be seen from FIG. 3, it is possible to change the speed of each gear by simply grasping any one of the clutches C1 to C4 and the brakes B1 and B2. Further, as shown in FIG. 3, the automatic transmission 10 according to the present embodiment can have a large speed ratio width and an appropriate speed ratio step.

  FIG. 4 is a diagram for explaining a schematic configuration of the drive device 6 and the like shown in FIG. 1 and a block diagram for explaining a main part of a control system provided in the vehicle for controlling the drive device 6 and the like. It is. The electronic control unit 90 includes a so-called microcomputer including a CPU, a ROM, a RAM, an input / output interface, and the like, and performs signal processing according to a program stored in advance in the ROM while using a temporary storage function of the RAM. Is basically executed, for example, output control of the engine 8 and shift control for automatically switching the gear stage of the automatic transmission 10 are performed. It is configured separately for use.

In FIG. 4, the operation amount Acc of the accelerator pedal 50 is detected by the accelerator operation amount sensor 52. The accelerator pedal 50 is largely depressed according to the driver's requested output amount, and corresponds to an accelerator operation member, and the accelerator operation amount Acc corresponds to the requested output amount. The intake pipe 53 of the engine 8 is provided with an electronic throttle valve 56 whose opening angle (opening) θ _TH is controlled by a throttle actuator 54.

Also, the fully closed state (idle state of the intake air quantity sensor 60, the electronic throttle valve 56 for detecting an intake air quantity Q of the engine rotational speed sensor 58, the engine 8 for detecting the rotational speed N _E of the engine 8 ) And a throttle valve opening sensor 62 with an idle switch for detecting the opening θ _TH , a vehicle speed sensor 64 for detecting the vehicle speed V (corresponding to the rotational speed N _OUT of the output shaft 28), a first motor generator MG1. MG1 rotational speed sensor 66 for detecting the rotational speed N _MG1 (= engine rotational speed N _E ), rotational speed N _MG2 of the second motor generator _MG2 (= rotational speed N _IN of the input shaft 16) MG2 rotational speed sensor 68, foot brake switch 70 for detecting that the brake pedal is operated, shift operation Operating position of a shift lever 72 provided in the work apparatus 71 (operation _position) the storage amount of _P lever position sensor 74 _{for SH} detecting the motor generators MG1, MG2 power storage device 76 connected to the (inverter 106) (residue (Capacity, charge amount) SOC sensor 78 for detecting SOC, MG1 temperature sensor 80 for detecting temperature T _MG1 of first motor generator MG1, MG2 temperature for detecting temperature T _MG2 of second motor generator MG2 A control current I _MG1 for controlling the first motor generator MG1 which is a generated current to the inverter 106 generated by the power generation of the sensor 81 or the first motor generator MG1 or a drive current from the inverter 106 for driving the first motor generator MG1. MG1 control current sensor for detection , The MG2 control current sensor 83 for detecting the control current I _MG2 for controlling the second motor generator MG2, which is the drive current from the inverter 106 for driving the second motor generator MG2, and the cooling of the engine 8 engine coolant temperature sensor 85 for detecting the water temperature I _W, oil temperature sensor 86 for detecting a temperature T _oIL of the hydraulic oil 46 in the automatic transmission 10, for detecting the temperature of the catalyst T _RE for purifying exhaust gas A catalyst temperature sensor 87, an acceleration sensor 88 for detecting the acceleration G of the vehicle, and the like are provided. From these sensors and switches, the engine speed N _E , the intake air amount Q, the throttle valve opening θ _TH , the vehicle speed V, first motor-generator rotational speed _{N MG1,} the second motor-generator rotational speed _{N MG2,} the foot brake Signal _{B ON} representing the operation of the presence or absence of work i.e. brake pedal 69, the operation position of the shift lever 72 _{P SH,} the remaining capacity SOC, first motor generator temperature _{T MG1,} second motor generator temperature _{T MG2,} the first motor-generator control current Signals representing I _MG1 , second motor generator control current I _MG2 , cooling water temperature I _W , oil temperature T _OIL , catalyst temperature T _RE , vehicle acceleration G, and the like are supplied to the electronic control unit 90.

Further, the electronic control device 90 receives a control signal for controlling the engine output, for example, a drive signal to the throttle actuator 54 for operating the opening degree θ _TH of the electronic throttle valve 56 and the fuel to the engine 8 by the fuel injection device 92. A fuel supply amount signal for controlling the supply amount, an ignition signal for instructing the ignition timing of the engine 8 by the ignition device 94, a supercharging pressure adjustment signal for adjusting the supercharging pressure, and an electric air conditioner driving signal for operating the electric air conditioner The control signal for controlling the inverter 106 by the MG1 controller 102 or the MG2 controller 104 for operating the first motor generator MG1 or the second motor generator MG2, power generation control (regeneration) control, etc., for operating the shift indicator Shift position (operation position) display signal, wheel of braking An ABS operation signal for operating an ABS actuator for preventing a lip, an AT shift solenoid 99 in a hydraulic control circuit 98 for operating the hydraulic actuators of the clutches C1 to C4 and brakes B1 and B2 of the automatic transmission 10, for example, linear Control of excitation and de-energization of the start clutch control valve 96 in the hydraulic control circuit 98 to operate the valve command signal for controlling the excitation and de-excitation of the solenoid valves SL1 to SL6 and the hydraulic actuator of the start clutch Ci. valve command signals for the linear solenoid valves of the hydraulic control circuit 98 for switching the lubricating flow F _J in other words lubricating pressure P _J of the hydraulic oil 46 in the lubricating fluid is circulated and supplied to each part of the drive apparatus 6 Valve command signal for controlling SLJ, said hydraulic friction engagement device A drive command signal (see FIG. 5) for operating an electric hydraulic pump 100 that is a hydraulic source of a hydraulic control circuit 98 that supplies hydraulic oil 46 for engaging the device or lubricating each part in the drive device 6; A signal for driving the electric heater, a control signal to a signal to the cruise control control computer, and the like are output.

FIG. 5 is a diagram schematically illustrating the configuration of the hydraulic control circuit 98 shown in FIG. In addition to the linear solenoid valves SL1 to SL6, the starting clutch control valve 96, the linear solenoid valve SLJ, and the electric hydraulic pump 100, the hydraulic control circuit 98 is sucked up by the electric hydraulic pump 100 and linear solenoid valves SL1 to SL6. And a relief valve type line pressure regulating valve 40 for regulating the hydraulic oil 46 in the oil pan 101 supplied to the starting clutch control valve 96 to the line hydraulic pressure PL, and a line for controlling the line hydraulic pressure PL according to the engine load and the like. pressure regulating valve 40 to control pressure P _L and supplies the linear solenoid valve SLT, and the oil cooler 44 to the hydraulic oil 46 discharged from the line pressure regulating valve 40 to cool the hydraulic oil 46 is supplied is provided.

  The hydraulic actuators (hydraulic cylinders) 34... 36, 38 of the clutches C1 to C4, the brakes B1 and B2, and the starting clutch Ci are respectively connected to the linear solenoid valves SL1 to SL using the line hydraulic pressure PL as a source pressure. The operating oil 46 adjusted by SL6 and the starting clutch control valve 96 is supplied, and the operating states of engagement / release of the clutches C1 to C4, the brakes B1 and B2, and the starting clutch Ci are changed. The line oil pressure PL is the maximum engagement pressure used to engage the clutches C1 to C4, the brakes B1 and B2, and the start clutch Ci. The linear solenoid valves SL1 to SL6 and the starting clutch control valve 96 have basically the same configuration, and are excited and de-energized independently by the electronic control unit 90, and the hydraulic actuators 34,. The hydraulic pressure of the hydraulic oil supplied to 38 is independently regulated.

Further, each part in the driving device 6, for example, each engaging part of the clutches C1 to C4, the brakes B1 and B2 and the starting clutch Ci, the first motor generator MG1, the second motor generator MG2, and the crankshaft 9 and the input shaft 16 etc. each bearings supporting the like output shaft 28, the line pressure regulating valve 40 hydraulic fluid 46 pressure regulated to a lubricating pressure P _J has been supplied by the discharge has been linear solenoid valve SLJ than, each unit in the driving device 6 Lubrication and cooling are performed. In this way, the hydraulic control circuit 98 circulates and supplies the hydraulic oil 46 as a lubricating fluid to each part in the drive device 6 and also cools the first motor generator MG1 and the second motor generator MG2 with the hydraulic oil 46. Functions as a circulation supply device.

FIG. 6 shows an example of the shift operation device 71. For example, the shift operation device 71 is disposed on the floor portion near the driver's seat, specifically, on the center console portion on the left side of the driver's seat. Further, the shift lever 72 is in the shift position P _SH having a shift pattern or a shift operation device 71 shown in FIG. 6, "P (parking)", "R (reverse)", "N (Neutral)", "D ( Drive) ”and“ M (manual) ”are operated according to the operation positions.

  The “P” position is a parking position. For example, when the manual valve in the hydraulic control circuit 98 is mechanically switched in accordance with the movement operation of the shift lever 72, the automatic transmission 10 is operated by the clutches C1 to C4 and the brakes B1 and B2. All of them are released and the power transmission is cut off, and the output shaft 28, that is, the drive wheels are mechanically fixed to be non-rotatable by a parking lock mechanism or the like according to the movement operation of the shift lever 72, for example. The “R” position is a reverse travel position where the reverse travel is performed. For example, the manual valve is mechanically switched in accordance with the movement operation of the shift lever 72, whereby a reverse gear stage can be established. The reverse gear stage “Rev1” or “Rev2” is established electrically according to the combined operation table. The “N” position is a power transmission cutoff position. For example, similarly to the “P” position, the automatic transmission 10 is brought into a power transmission cutoff state by releasing all of the clutches C1 to C4 and the brakes B1 and B2.

  The “D” position is a forward travel position in which the forward gear of the automatic transmission 10 is automatically switched and travels forward, that is, a forward travel position. For example, the manual valve is mechanically switched according to the movement operation of the shift lever 72. Thus, it is possible to establish all the forward gear stages “1st” to “8th”, and the forward gear stages “1st” to “8th” are electrically established according to the engagement operation table of FIG. . That is, when the shift lever 72 is operated to the “D” position, a D range (full range automatic transmission mode) is established in which the gears are automatically shifted using the forward gears “1st” to “8th”.

  The “M” position is a forward manual shift travel position that establishes a manual shift mode in which a forward shift is established by manually switching the shift range or gear position of the forward gear of the automatic transmission 10, that is, the forward manual shift travel position. For example, when the shift lever 72 is operated to the “M” position, the manual shift mode is established. The “+” position provided in the “M” position is an operation position for shifting the shift range or the gear stage up for each operation in the manual shift mode. Similarly, the “−” position is an operation position for shifting the shift range or the gear stage to the down side for each operation in the manual shift mode.

FIG. 7 is a functional block diagram for explaining a main part of the control function provided in the electronic control unit 90. In FIG. 7, the engine output control means 110 controls the fuel injection amount by the fuel injection device 92 for controlling the fuel injection amount, in addition to controlling the opening and closing of the electronic throttle valve 56 by the throttle actuator 54 for the throttle control. For the timing control, the output control of the engine 8 is executed by controlling the ignition timing by an ignition device 94 such as an igniter. For example, the engine output control means 110 drives the throttle actuator 54 based on the actual accelerator operation amount Acc from the pre-stored relationship shown in FIG. 8, and increases the throttle valve opening θ _TH as the accelerator operation amount Acc increases. The throttle control is executed so as to increase.

The shift control means 112 is based on the actual vehicle speed V and the throttle valve opening θ _TH from the relationship (shift map, shift diagram) stored in advance with the vehicle speed V and the throttle valve opening θ _TH shown in FIG. 9 as parameters, for example. The shift stage (gear stage) to be switched by the automatic transmission 10 is determined, and a shift command (shift output) is transmitted to the hydraulic control circuit 98 based on, for example, the engagement operation table of FIG. 3 so as to obtain the determined shift stage. Is output to execute the shift control for automatically switching the shift stage of the automatic transmission 16.

  The hydraulic control circuit 98 executes excitation, de-excitation, and current control of the linear solenoid valves SL1 to SL6 in accordance with a shift command from the shift control means 112, and engages and disengages clutches C1 to C4 and brakes B1 and B2. To establish one of the first shift speed “1st” to the eighth shift speed “8th”, or one of the reverse gears “Rev1” and “Rev2”. Control hydraulic pressure. For example, as shown in the engagement operation table of FIG. 3, in the upshift from the 4th gear to the 5th gear, the clutch C4 disengagement transient hydraulic pressure and the clutch C2 are engaged so that the clutch C4 is disengaged and the clutch C2 is engaged. The combined transient hydraulic pressure is controlled. It should be noted that various modes are possible, such as performing shift control based on the accelerator opening Acc, the intake air amount Q, the road gradient, and the like.

For example, the shift line in the shift diagram of FIG. 9 should execute a shift on the shift line whether or not the actual vehicle speed V crosses the line on the horizontal line indicating the actual throttle valve opening θ _TH (%). the value is for determining whether exceeds (shift point vehicle speed) V _S, is stored in advance as a series of the values V _S that shift point vehicle speed. For example, as the vehicle speed V increases or the throttle valve opening _θTH decreases, a high-speed gear stage with a small gear ratio is established. In the shift diagram of FIG. 9, the down shift line is omitted. Further, the accelerator operation amount Acc (%) may be used in place of the throttle valve opening θ _TH (%).

  The hybrid control means 114 is for performing traveling in a plurality of operation modes in which the operating states of the engine 8 and the motor generators MG1 and MG2 such as motor traveling, engine traveling, motor and engine traveling differ according to the traveling state of the vehicle. Then, opening / closing control of the starting clutch Ci, power running control of the first motor generator MG1 and second motor generator MG2, regeneration control, and the like are executed. FIG. 10 shows an example of the operation mode that is normally performed.

In FIG. 10, in the engine travel mode in which the vehicle travels exclusively using the engine 8 as a driving force source for travel, for example, even when the remaining capacity SOC of the power storage device 76 decreases, the vehicle travels or is compared with motor travel. Therefore, the hybrid control means 114 outputs a command to the hydraulic control circuit 98 so that the start clutch Ci is engaged by the start clutch control valve 96 for traveling that requires a greater driving force. Is directly transmitted to the input shaft 16 of the automatic transmission 10, and a command is output to the engine output control means 110 so that the engine 8 generates a necessary driving force and travels. Further, in a case the remaining capacity SOC of power storage device 76 is small, the first motor generator MG1 is a generator (regeneration) state if necessary, MG1 controller as its power energy E _D is in the power storage device 76 A command is output to 102.

In the motor start / run mode in which the vehicle starts / runs exclusively using the second motor generator MG2 as a driving force source for travel, the hybrid control means 114 performs start clutch control for quiet vehicle start and travel, for example. A command is output to the hydraulic control circuit 98 so that the starting clutch Ci is released by the valve 96 so that the power transmission path between the engine 8 and the automatic transmission 10 is cut off, and a drive current is supplied from the inverter 106. Then, the second motor generator MG2 is set in a power running state, and a command is output to the MG2 controller 104 so that the second motor generator MG2 generates a necessary driving force and travels. At this time, since the power transmission path between the engine 8 and the automatic transmission 10 is cut off, a reduction in fuel consumption due to dragging of the engine 8 that is not operating is suppressed. Further, when the remaining capacity SOC of the power storage device 76 is small, a command is output to the engine output control means 110 to operate the engine 8 as necessary, and the first motor generator MG1 is set in a power generation state. generating energy _{E D} and outputs a command to the MG1 controller 102 as in the power storage device 76. Further, it may be directly supplied its power energy E _D is as the drive current of the second motor generator MG2 through the inverter 106.

  In addition, the start clutch Ci that has been released in the motor start mode is brought into the slip engagement state, and in addition to the output of the second motor generator MG2, a part of the output of the engine 8 is transmitted to the drive wheels 32, which is necessary. The vehicle may be started by generating a driving force. That is, the hybrid control unit 114 outputs a command to the engine output control unit 110 so as to operate the engine 8 in addition to setting the second motor generator MG2 in the power running state, and the start clutch Ci is controlled by the start clutch control valve 96. A command is output to the hydraulic control circuit 98 so that is slip-engaged (semi-engaged), and the second motor generator MG2 and the engine 8 generate necessary driving force to start the vehicle. At this time, the hybrid control means 114 outputs a command to the hydraulic control circuit 98 so that the start clutch Ci is completely engaged by the start clutch control valve 96, for example, when the vehicle speed V becomes equal to or higher than a predetermined vehicle speed. In the case of starting the motor by adding a part of the output of the engine 8 as described above, compared with the case where the necessary driving force is obtained exclusively by the output of the second motor generator MG2, the amount of driving force generated by the output of the engine 8 is reduced. Therefore, the output of second motor generator MG2 is reduced. Further, the motor start with a part of the output of the engine 8 is normally performed in the motor start mode.

  In the engine + motor traveling mode in which the vehicle travels using the engine 8 and the second motor generator MG2 as a driving force source for traveling, for example, for acceleration traveling, the hybrid control means 114 is started by the start clutch control valve 96. A command is output to the hydraulic control circuit 98 so that the clutch Ci is engaged, and the output of the engine 8 is directly transmitted to the input shaft 16 of the automatic transmission 10, and a necessary driving force is generated by the engine 8. A command is output to the engine output control means 110 so that the vehicle travels, and a driving current is supplied from the inverter 106 so that the second motor generator MG2 is in a power running state, and a necessary driving force is generated by the second motor generator MG2. A command is output to the MG2 controller 104 so as to travel. Further, a drive current may be supplied from inverter 106 to cause first motor generator MG1 to be in a power running state, and a command may be output to MG1 controller 102 so that the first motor generator MG1 generates a drive force and travels.

  In addition to the normal operation mode shown in FIG. 10, for example, an electric system failure (failure) including the first motor generator MG1 and the second motor generator MG2, or a decrease in the output of the second motor generator MG2, etc. In the case where the electrical system function is deteriorated due to such an extremely low temperature, the hybrid control means 114 outputs a command to the engine output control means 110 so as to operate the engine 8 and the start clutch control valve 96 starts the operation. A command is output to the hydraulic control circuit 98 so that the clutch Ci is slip-engaged (half-engaged), and a so-called friction start is executed in which the vehicle is started while the necessary driving force is gradually generated only by the engine 8. At this time, the hybrid control means 114 outputs a command to the hydraulic control circuit 98 so that the start clutch Ci is completely engaged by the start clutch control valve 96, for example, when the vehicle speed V becomes equal to or higher than a predetermined vehicle speed.

  The friction start determination unit 116 determines whether or not to execute the friction start. For example, the friction start determination unit 116 determines the execution of the friction start when an electric system failure (failure) including the first motor generator MG1 and the second motor generator MG2 occurs. Alternatively, the friction start determining means 116 is such that the temperature of the power storage device 77, the second motor generator MG2, etc. is equal to or lower than a predetermined temperature that has been experimentally determined in advance to cause a decrease in the function of the electric system. In this case, the execution of the friction start is determined. The hybrid control unit 114 switches the driving mode when the vehicle starts based on the determination result of the friction start determination unit 116.

Lubricating pressure control means 118 controls the lubricating pressure P _J of the hydraulic oil 46 is circulated and supplied to the driving unit 6 by the hydraulic control circuit 98. Specifically, the lubrication pressure controller 118 outputs a valve command signal to the hydraulic control circuit 98 so as to control the lubricant pressure P _J of the hydraulic oil 46 by the linear solenoid valve SLJ. For example, the lubricating pressure P _J is increased so that the flow rate of the hydraulic oil 46 increases as the first motor generator temperature T _MG1 and the second motor generator temperature T _MG2 become relatively higher than when the temperature is low. Based on the first motor generator temperature T _MG1 and the second motor generator temperature T _MG2 , experimentally obtained and set in advance so that the cooling performance of the first motor generator MG1 and / or the second motor generator MG2 is improved. Yes. That is, the lubricating pressure control means 118, by controlling the lubricating pressure P _J of the hydraulic oil 46, which functions as a lubricating flow rate control means for controlling the lubrication flow F _J of the hydraulic oil 46 is circulated and supplied to the driving device 6 .

By the way, when the friction start is executed, the first motor generator MG1 and the second motor generator MG2 are not in an operating state, so that the temperature rise is suppressed. Even in the motor start mode, when the start clutch Ci is in the slip engagement state and the vehicle is started using a part of the output of the engine 8 as a driving force in addition to the output of the second motor generator MG2, the engine only 8 driving force caused by the output of the output of the second motor generator MG2 is small, the amount of heat generated or the temperature rise is suppressed by reduction of the second motor generator MG2, the are small power energy E _D 1 Temperature rise of motor generator MG1 is suppressed.

Therefore, when the lubricating pressure P _J is uniformly set on the basis of the first motor generator temperature T _MG1 and second motor generator temperature T _MG2 in this case, operation required for cooling the starting clutch Ci and hydraulic oil 46 There was a possibility that the oil 46 would not flow.

Therefore, the lubricating pressure control means 118 is driven by the hydraulic control circuit 98 based on the torque capacity of the starting clutch Ci when the engine 8 is operated so that the flow rate of the hydraulic oil 46 necessary for cooling can be obtained appropriately. the lubricating pressure P _J of the hydraulic oil 46 is circulated and supplied to the device 6, ie the lubricating flow F _J of the hydraulic control circuit 98 by the hydraulic fluid 46 to be circulated and supplied to the driving device 6, and controls. Regard determination of the lubricating pressure P _J, will be described below, for the lubrication flow F _J is the same as the lubrication pressure P _J omitted.

The motor temperature state monitoring means 120 is configured to detect the temperature state of the first motor generator MG1 and / or the second motor generator MG2, for example, the first motor generator temperature T _MG1 , the second motor generator temperature T _MG2 , the first motor generator temperature change T _MG1 ′. (= DT _MG1 / dt) and second motor generator temperature change T _MG2 ′ (= dT _MG2 / dt) are sequentially monitored.

The motor heat generation state monitoring means 122 sequentially monitors the heat generation state of the first motor generator MG1 and / or the second motor generator MG2, for example, the heat generation amount or the heat generation amount change. For example, the motor heat generation state monitoring means 122 generates the heat generation amount Q _MG1 for a predetermined time, for example, 1 second of the first motor generator MG1 based on the first motor generator control current I _MG1 and the predetermined voltage V _{MG1 of} the first motor generator _MG1. (= I _MG1 * V _MG1 * 1) is sequentially calculated and monitored. Similarly, the motor heat generation state monitoring unit 122 generates the heat generation amount Q for the second motor generator MG2 for a predetermined time, for example, 1 second, based on the second motor generator control current _IMG2 and the predetermined voltage _{VMG2 of} the second motor generator MG2. _MG2 (= I _MG2 * V _MG2 * 1) is sequentially calculated and monitored. In addition, the motor heat generation state monitoring means 122 performs the first motor generator heat generation amount change Q _MG1 ′ (= dQ _MG1 / dt) based on the first motor generator heat generation amount Q _MG1 and the second motor generator heat generation amount Q _MG2. Second motor generator heat generation amount change Q _MG2 ′ (= dQ _MG2 / dt) is sequentially monitored.

The hydraulic oil temperature state monitoring means 124 sequentially monitors the temperature state of the hydraulic oil 46, for example, the oil temperature T _OIL and the oil temperature change T _OIL ′ (= dT _OIL / dt).

Clutch engagement rate correction factor calculation unit 126 is subjected to the lubricating pressure P _J in order to correct the lubricating pressure P _J based on the torque capacity of the starting clutch Ci during vehicle starting when the engine operation is usually carried out A correction coefficient K _Ci is calculated.

When the start clutch Ci is slip-engaged at the start of the motor and the output (power) of the engine 8 is transmitted to the drive wheels 32 according to the torque capacity of the start clutch Ci, the second motor is used to obtain the same driving force. Since the output of the generator MG2 is reduced by the amount corresponding to the driving force generated by the output of the engine 8, the heat generation amount and the heat generation amount change of the second motor generator MG2 are relatively reduced. When the output of the second motor generator MG2 is relatively small, since also small power energy E _D by the first motor generator MG1, the heating value or heat value change of the first motor generator MG1 is also relatively It is made smaller. Then, since the possibility that the first motor generator temperature T _MG1 and / or the second motor generator temperature T _MG2 becomes high becomes relatively low, it is necessary for cooling the first motor generator MG1 and / or the second motor generator MG2. The flow rate of the hydraulic oil 46 may be relatively small. Therefore, the clutch engagement rate correction factor calculation unit 126, as the torque capacity of the starting clutch Ci is high, i.e. greater the output of the engine 8 is transmitted to the drive wheel 32 is increased, lowering the lubrication pressure P _J actuation The correction coefficient K _Ci is reduced so that the flow rate of the oil 46 decreases.

From another viewpoint, if the start clutch Ci is slip-engaged and the heat generation amount and the heat generation amount change of the first motor generator MG1 and the second motor generator MG2 are relatively reduced, the hydraulic oil 46 may be warmed. Therefore, the clutch engagement rate correction coefficient calculating means 126 decreases the correction coefficient K _Ci as the torque capacity of the starting clutch Ci increases. Alternatively, the higher the torque capacity of the starting clutch Ci, the smaller the slip amount of the starting clutch Ci and the lower the amount of heat generated by the starting clutch Ci itself. Therefore, the flow rate of the hydraulic oil 46 necessary for cooling the starting clutch Ci Since the possibility that the hydraulic oil 46 is warmed is relatively low, the clutch engagement rate correction coefficient calculating means 126 increases the correction coefficient as the torque capacity of the starting clutch Ci increases. _Reduce K _Ci .

On the contrary, since the output (power) of the engine 8 transmitted to the drive wheels 32 becomes relatively smaller as the torque capacity of the starting clutch Ci at the time of starting the vehicle when the engine is activated, the same driving force is obtained. In order to obtain this, the output of the second motor generator MG2 is relatively increased, and the amount of generated heat and change in the amount of generated heat are relatively increased. Similarly, when the output of the second motor generator MG2 is relatively large, since it is also larger power energy E _D by the first motor-generator MG1, also relative calorific or heating value change of the first motor generator MG1 To be bigger. Then, since it is predicted that the first motor generator temperature T _MG1 and / or the second motor generator temperature T _MG2 will be relatively high, it is necessary for cooling the first motor generator MG1 and / or the second motor generator MG2. The flow rate of the hydraulic oil 46 is also relatively increased. Therefore, the clutch engagement rate correction factor calculation means 126, in order to enhance the cooling effect, as the torque capacity of the starting clutch Ci is low, the correction so that the flow rate of the hydraulic fluid 46 by increasing the lubricant pressure P _J is increased The coefficient K _Ci is increased. In other words, the torque capacity of the starting clutch Ci is low even when the first motor generator temperature T _MG1 and / or the second motor generator temperature T _MG2 is not so high that the lubricating pressure P _J has to be increased. Therefore, it is predicted that the first motor generator temperature T _MG1 and / or the second motor generator temperature T _MG2 will rise. Therefore, the clutch engagement rate correction coefficient calculating means 126 may increase the starting clutch Ci in order to enhance the cooling effect. The correction coefficient K _Ci is increased so that the lubricating pressure P _J is increased and the flow rate of the hydraulic oil 46 is increased as the torque capacity of the hydraulic oil 46 becomes lower.

Alternatively, the lower the torque capacity of the starting clutch Ci when the engine is operating, the larger the slip amount of the starting clutch Ci, and the higher the heat generation amount, the higher the temperature of the starting clutch Ci may be. since the clutch engagement rate correction factor calculation means 126, in order to enhance the cooling effect of the starting clutch Ci, higher the torque capacity of the starting clutch Ci is low, the flow rate of the hydraulic oil 46 is increased by increasing the lubricant pressure P _J Thus, the correction coefficient K _Ci is increased.

In this embodiment, the actual torque capacity of the start clutch Ci is a ratio of the actual torque capacity to the maximum torque capacity when the start clutch Ci is completely engaged, that is, a start clutch that is 1.0 when it is the maximum torque capacity. It is represented by the engagement rate C _{Ci of Ci} (= actual torque capacity / maximum torque capacity). For example, the relationship (map) between the engagement rate C _Ci of the starting clutch Ci and the hydraulic pressure command value to the hydraulic control circuit 98 for engaging the starting clutch Ci is determined experimentally in advance. Then, the clutch engagement rate correction coefficient calculating unit 126 calculates the engagement rate C _Ci of the starting clutch Ci based on the hydraulic pressure command value from the hybrid control unit 114 based on the predetermined relationship.

FIG. 11 is an example of a relationship (map) that is experimentally determined and determined in advance between the engagement rate C _Ci of the start clutch _Ci and the correction coefficient K _Ci when the engine is activated during normal vehicle start. Yes, the correction coefficient K _Ci is made smaller as the engagement rate (torque capacity) C _{Ci of} the starting clutch Ci becomes higher.

The clutch slip ratio calculating means 128 determines the slip amount (= maximum torque capacity−torque capacity) of the starting clutch Ci at the friction start as a ratio of the actual slip amount to the maximum slip amount when the starting clutch Ci is completely released, that is, the maximum slip amount. Is the slip ratio C _SP (= actual slip amount / maximum slip amount) of the starting clutch Ci, that is, the maximum torque capacity and the actual torque with respect to the maximum torque capacity when the starting clutch Ci is completely engaged. The ratio of the difference from the capacity is calculated as (= 1.0− (actual torque capacity / maximum torque capacity)). For example, the relationship between the hydraulic pressure command value to the hydraulic control circuit 98 for engaging the slip ratio C _SP and starting clutch Ci of the starting clutch Ci (map) is determined experimentally determined in advance. The clutch slip ratio calculating unit 128 calculates the slip ratio C _SP of the starting clutch Ci based on the hydraulic pressure command value by the hybrid control means 114 from the predetermined relationship. Further, the engagement ratio _{C Ci} slip ratio _{C SP} and starting clutch Ci of the starting clutch Ci, a relation of complement (extra number) for 1.0 to one another.

When the friction start determining unit 116 determines that the friction start is not to be executed, the lubrication pressure determining unit 130 and the first motor generator MG1 and / or the first motor generator MG1 that are sequentially monitored by the motor temperature state monitoring unit 120 are used. 2 The temperature state of the motor generator MG2 and / or the heat generation state of the first motor generator MG1 and / or the second motor generator MG2 and / or the hydraulic oil temperature state monitoring which are sequentially monitored by the motor heat generation state monitoring means 122 Based on the temperature state of the hydraulic oil 46 that is sequentially monitored by the means 124 and the correction coefficient K _Ci calculated by the clutch engagement rate correction coefficient calculation means 126, the lubrication pressure is determined when the vehicle is normally started. Controlled by control means 118 Determining the lubricant pressure _{P J.}

For example, the lubricating pressure determining means 130 determines the lubrication pressure _{P J} on the basis of the first motor generator temperature _{T MG1} or the second motor-generator temperature _{T MG2.} Specifically, the lubrication pressure determining means 130 needs to cool the first motor generator MG1 and / or the second motor generator MG2 itself as the first motor generator temperature T _MG1 or the second motor generator temperature T _MG2 is higher. There relatively high for, increasing the lubricant pressure P _J so that the flow rate of the hydraulic oil 46 is increased to the cooling performance is increased. From another point of view, the higher the first motor generator temperature T _MG1 or the second motor generator temperature T _MG2 is, the higher the possibility that the hydraulic oil 46 is warmed. increase the pressure _{P J.}

FIG. 12 is an example of a relationship (map) that has been experimentally determined and determined in advance between the first motor generator temperature T _MG1 or the second motor generator temperature T _MG2 and the lubricating pressure P _{J. MG1} or higher the second motor generator temperature _{T MG2} increases, lubricating pressure _{P J} has been increased.

The lubricating pressure determining means 130 determines the lubricating pressure P _J based on the first motor generator heat generation amount Q _MG1 or the second motor generator heat generation amount Q _MG2 . Specifically, the lubrication pressure determining means 130 increases the first motor generator temperature T _MG1 and / or the second motor generator temperature T _{MG2 as} the first motor generator heat generation amount Q _MG1 or the second motor generator heat generation amount Q _MG2 is higher. because potential increases relatively high, increasing the lubricant pressure P _J so that the flow rate of the hydraulic oil 46 is increased to the cooling performance is increased. From another point of view, the higher the first motor generator calorific value Q _MG1 or the second motor generator calorific value Q _MG2 is, the higher the possibility that the hydraulic oil 46 is warmed. increase the lubricant pressure _{P J} is.

FIG. 13 is an example of a relationship (map) that has been experimentally determined and determined in advance between the first motor generator heat generation amount Q _MG1 or the second motor generator heat generation amount Q _MG2 and the lubricating pressure P _J. The higher the heat generation amount Q _MG1 or the second motor generator heat generation amount Q _MG2 is, the larger the lubricating pressure P _J is.

Further, the lubricating pressure determining means 130 determines the lubricating pressure P _J based on the first motor generator temperature change T _MG1 ′ or the second motor generator temperature change T _MG2 ′. Specifically, the lubricating pressure determination means 130 determines that the first motor generator temperature T _MG1 and / or the second motor generator temperature increases as the first motor generator temperature change T _MG1 ′ or the second motor generator temperature change T _MG2 ′ increases. Since the possibility that T _MG2 becomes high is relatively high, the lubricating pressure P _J is increased so that the flow rate of the hydraulic oil 46 is increased and the cooling performance is improved. From another viewpoint, the greater the first motor generator temperature change T _MG1 ′ or the second motor generator temperature change T _MG2 ′, the higher the possibility that the hydraulic oil 46 is warmed. means 130 raises the lubricating pressure _{P J.}

FIG. 14 is an example of a relationship (map) determined in advance experimentally between the first motor generator temperature change T _MG1 ′ or the second motor generator temperature change T _MG2 ′ and the lubricating pressure P _J. As the motor generator temperature change T _MG1 ′ or the second motor generator temperature change T _MG2 ′ becomes higher, the lubricating pressure P _J is increased.

The lubricating pressure determining means 130 determines the lubricating pressure P _J based on the first motor generator heat generation amount change Q _MG1 ′ or the second motor generator heat generation amount change Q _MG2 ′. Specifically, the lubrication pressure determination means 130 determines that the first motor generator temperature T _MG1 and / or the second motor increases as the first motor generator heat generation amount change Q _MG1 ′ or the second motor generator heat generation amount change Q _MG2 ′ increases. Since the possibility that the generator temperature T _MG2 becomes high is relatively high, the lubricating pressure P _J is increased so that the flow rate of the hydraulic oil 46 is increased and the cooling performance is improved. From another viewpoint, the greater the first motor generator heat generation amount change Q _MG1 ′ or the second motor generator heat generation amount change Q _MG2 ′, the higher the possibility that the hydraulic oil 46 is warmed. pressure determining means 130 raises the lubricating pressure _{P J.}

FIG. 15 is an example of a relationship (map) that has been experimentally determined and determined in advance between the first motor generator heat generation amount change Q _MG1 ′ or the second motor generator heat generation amount change Q _MG2 ′ and the lubrication pressure P _J. As the first motor generator heat generation amount change Q _MG1 ′ or the second motor generator heat generation amount change Q _MG2 ′ increases, the lubricating pressure P _J is increased.

Further, the lubricating pressure determining means 130, the lubricant pressure P in order to correct the lubricating pressure P _J, based on the oil temperature T _OIL or oil temperature change T _{OIL 'of} the hydraulic fluid 46 during the vehicle start is usually carried out _A correction coefficient K _OIL to be multiplied by _J is determined. Specifically, the lubrication pressure determining means 130 has a relatively higher necessity to cool the hydraulic oil 46 or the warmed hydraulic oil 46 as the oil temperature T _OIL or the oil temperature change T _OIL ′ increases. In other words, it is predicted that the oil temperature T _OIL may be relatively high and the flow rate of the hydraulic oil 46 necessary for cooling the oil 46 needs to be relatively increased. the flow rate of the hydraulic oil 46 by increasing the lubricant pressure P _J has a larger correction coefficient K _oIL to increase.

FIG. 16 shows the relationship (map) determined and obtained experimentally in advance between the oil temperature T _OIL or the oil temperature change T _OIL ′ of the hydraulic oil 46 and the correction coefficient K _OIL when the vehicle is normally started. As an example, the correction coefficient K _OIL is increased as the oil temperature T _OIL or the oil temperature change T _OIL ′ of the hydraulic oil 46 increases.

Then, the lubricating pressure determining means 130 determines, for example, the lubricating pressure P _J determined based on the relationship shown in FIGS. 12 to 15 based on the temperature state and heat generation state of the first motor generator MG1 or the second motor generator MG2, for example. adding to determine the lubricating pressure P _J underlying. Further, the lubricating pressure determining means 130 multiplies the basic lubricating pressure P _JB by the correction coefficient K _Ci and the correction coefficient K _OIL, and the lubricating pressure control means 118 at the time of normal vehicle start. The target lubricating pressure P _J ^* (= P _JB * K _Ci * K _OIL ) controlled by is determined.

When the friction start determining means 116 determines that the friction start is to be executed, the lubricating pressure determining means 130 is the temperature state of the hydraulic oil 46 that is successively monitored by the hydraulic oil temperature state monitoring means 124. , and based on the slip ratio C _SP of the starting clutch Ci in the friction start calculated by the clutch slip ratio calculating unit 128, determines the lubricating pressure P _J, which is controlled by the lubricating pressure control means 118 when the friction start To do.

For example, the lubricating pressure determining means 130 determines the lubricating pressure P _J based on the oil temperature T _OIL or the oil temperature change T _OIL ′ of the hydraulic oil 46 at the time of friction start. Specifically, the lubricating pressure determining means 130 has a relatively high necessity for cooling the hydraulic oil 46 itself or the hydraulic oil 46 is warmed as the oil temperature T _OIL or the oil temperature change T _OIL ′ is larger. since relatively high probability, increasing the lubricant pressure P _J as to increase the flow rate of the hydraulic fluid 46 to improve cooling performance.

FIG. 17 is an example of a relationship (map) determined and obtained experimentally in advance between the oil temperature T _OIL or oil temperature change T _OIL ′ of the hydraulic oil 46 at the time of friction start and the lubricating pressure P _J. the oil temperature _{T oIL} or oil temperature change in the oil 46 _{T oIL} 'enough to increase the lubricating pressure _{P J} has been increased.

Further, the lubricating pressure determining means 130 determines the correction factor K _SP that are subjected to lubricating pressure P _J in order to correct the lubricating pressure P _J based on the slip ratio C _SP of the starting clutch Ci during friction start To do. Specifically, the lubrication pressure determining means 130, the higher the slip ratio C _SP of the starting clutch Ci, the relative likelihood that the temperature rises in the heating value is relatively large is by starting clutch Ci of the starting clutch Ci In other words, since the heat generation amount of the start clutch Ci is relatively large and the temperature of the start clutch Ci is predicted to be relatively high, the hydraulic oil 46 necessary for cooling the start clutch Ci the flow rate must also be relatively increased, to increase the correction coefficient K _SP so that the flow rate of the hydraulic fluid 46 by increasing the lubricant pressure P _J increases. From another point of view, when the heat generation amount of the start clutch Ci is relatively large and the temperature of the start clutch Ci is increased, the possibility that the hydraulic oil 46 is warmed is relatively high. It is, as the slip ratio _{C SP} of the starting clutch Ci is increased, increasing the correction factor _{K SP.}

FIG. 18 is an example of a relationship (map) determined experimentally in advance between the slip ratio C _SP of the start clutch Ci at the time of friction start and the correction coefficient K _SP, and the slip ratio C of the start clutch Ci. _The correction coefficient _KSP is increased as the _SP increases.

Then, for example, the lubricating pressure determining means 130 _applies the lubricating pressure P _J determined based on the oil temperature T _OIL or the oil temperature change T _OIL ′ of the hydraulic oil 46 at the time of friction start from the relationship shown in FIG. The correction coefficient K _SP is multiplied to determine a target lubricating pressure P _J ^* (= P _J * K _SP ) controlled by the lubricating pressure control means 118 at the time of friction start. Thus, when the friction start, although the lubricating pressure P _J is determined based on the slip ratio C _SP of the starting clutch Ci, engagement of the slip ratio C _SP and starting clutch Ci of the starting clutch Ci as described above Since the joint ratio C _Ci has a complement (remainder) relationship with respect to 1.0, the lubricating pressure P _J is determined based on the engagement ratio (torque capacity) C _Ci of the start clutch Ci at the time of friction start. It can be said that.

The lubricating pressure control means 118, to be switched lubricating pressure P _J of the hydraulic oil 46 in the lubricating pressure target P _J ^* by the linear solenoid valve SLJ, it outputs a valve command signal to the hydraulic control circuit 98.

19, the flow rate of the electronic control unit 90 for controlling operation of the main portion, that the vehicle is started during the drive of the hydraulic fluid 46 to be circulated and supplied to 6 lubrication pressure P _J hydraulic oil 46 necessary for the control to cool the It is a flowchart explaining the control operation | movement made to obtain appropriately, for example, is repeatedly performed by the very short cycle time of about several msec thru | or several tens msec.

  First, in step (hereinafter, step is omitted) S1 corresponding to the friction start determination means 116, it is determined whether or not to execute the friction start. For example, when a failure (failure) of the electric system including the first motor generator MG1 and the second motor generator MG2 occurs, or the temperature of the power storage device 77, the second motor generator MG2, and the like deteriorates the function of the electric system. The execution of the friction start is determined when the occurrence is below a predetermined temperature determined experimentally in advance.

When the determination in S1 is negative, the temperature state of the first motor generator MG1 and / or the second motor generator MG2 is determined in S2 corresponding to the motor temperature state monitoring unit 120 and / or the motor heat generation state monitoring unit 122. The heat generation state of the first motor generator MG1 and / or the second motor generator MG2 is monitored. Next, in S3 corresponding to the hydraulic oil temperature state monitoring means 124, the temperature state of the hydraulic oil 46 is monitored. Further, in the subsequent S4 corresponding to the clutch engagement rate correction coefficient calculating means 126, for example, the engagement of the start clutch Ci at the time of engine operation at the time of vehicle start normally performed from the relationship (map) shown in FIG. A correction coefficient K _Ci is calculated based on the rate C _Ci .

Next, in S5 corresponding to the lubricating pressure determining means 130, for example, the temperature state and heat generation state of the first motor generator MG1 or the second motor generator MG2 monitored in S2 from the relationships (maps) shown in FIGS. lubricating pressure P _J is determined based on. For example, the correction coefficient K _OIL is determined based on the oil temperature T _OIL or the oil temperature change T _OIL ′ of the hydraulic oil 46 monitored in S3 from the relationship (map) shown in FIG. Then, the basic lubricating pressure P _JB determined based on the temperature state and the heat generation state is multiplied by the correction coefficient K _OIL and the correction coefficient K _Ci calculated in S4, so that the vehicle start normally performed is performed. The target lubricating pressure P _J ^* (= P _JB * K _Ci * K _OIL ) is determined.

Alternatively, if the determination in S1 is affirmative, the temperature state of the hydraulic oil 46 is monitored in S7 corresponding to the hydraulic oil temperature state monitoring unit 124. Next, in S8 corresponding to the clutch slip ratio calculating means 128 in advance at the time of the friction start the hydraulic pressure command value to the hydraulic control circuit 98 for engaging the slip ratio C _SP and starting clutch Ci of the starting clutch Ci slip ratio C _SP of the starting clutch Ci based on the hydraulic pressure command value by the hybrid control means 114 is calculated from the experimentally determined relationship (map).

Next, in S5 corresponding to the lubricating pressure determining means 130, for example, based on the oil temperature T _OIL or the oil temperature change T _OIL ′ of the hydraulic oil 46 monitored in S7 from the relationship (map) shown in FIG. P _J is determined. The correction factor K _SP is determined, for example, from the relationship shown in FIG. 18 (map) based on the slip ratio C _SP of the starting clutch Ci, which is calculated in S8. Then, the lubricating pressure P _J determined based on the oil temperature T _OIL or the oil temperature change T _OIL ′ of the hydraulic oil 46 is multiplied by the correction coefficient K _SP to obtain a target lubricating pressure P at the time of friction start. _J ^* (= P _J * K _SP ) is determined.

Followed, wherein in S6 corresponding to the lubricating pressure control means 118, to be switched lubricating pressure P _J of the hydraulic oil 46 in the lubricating pressure target determined in the S5 P _J ^* by the linear solenoid valve SLJ, the hydraulic control circuit A valve command signal is output to 98.

As described above, according to this embodiment, lubricating pressure P _J of the hydraulic oil 46 is circulated and supplied to the driving unit 6 by the hydraulic control circuit 98, the starting clutch Ci during operation of the engine 8 engagement ratio (torque Since the lubricating pressure control means 118 is controlled based on the capacity) C _Ci , the cooling of the first motor generator MG1 and / or the second motor generator MG2 whose driving state changes according to the engagement rate C _Ci of the starting clutch Ci. The flow rate required for the above can be obtained appropriately. Thereby, the durability of first motor generator MG1 and / or second motor generator MG2 is improved. Further, deterioration of fuel efficiency for not increase constantly flow by increasing the lubricant pressure P _J uniformly is suppressed. In addition, the flow rate necessary for cooling the start clutch Ci whose calorific value changes according to the engagement rate C _Ci is appropriately obtained, and the durability of the start clutch Ci is improved.

For example, the lower the engagement rate C _Ci of the starting clutch Ci when starting the vehicle when the engine is operating, the smaller the output of the engine 8 transmitted to the drive wheels 32 is. In order to obtain the power), the output of the second motor generator MG2 is relatively increased, so that the electric energy (driving current) required for driving the second motor generator MG2 is increased and the first motor generator MG1 generates power. energy _{E D} is relatively large. Then, it is predicted that the heat generation amount (Q _MG1 , Q _MG2 ) of first motor generator MG1 and / or second motor generator MG2 will rise and their temperatures (T _MG1 , T _MG2 ) will become relatively high. since the lubricating pressure control means 118, as engagement ratio C _Ci of the starting clutch Ci during engine operation is lower, by increasing the lubricant pressure P _J to increase the flow rate of the hydraulic oil 46, increasing the cooling effect.

In other words, even if the temperature (T _MG1 , T _MG2 ) of the first motor generator MG1 and / or the second motor generator MG2 is not high enough to increase the lubricating pressure P _J , the starting clutch Ci When the engagement rate C _Ci of the engine is low, the temperature (T _MG1 , T _MG2 ) is expected to increase, so that the lubricating pressure P _J is increased to increase the flow rate of the hydraulic oil 46 and the cooling effect is increased. Can be increased.

Alternatively, as the engagement ratio C _Ci of the starting clutch Ci during engine operation is lower, the slip ratio C _SP of the starting clutch Ci is increased, the temperature of the relatively starting clutch Ci its calorific value is increased is higher there is a possibility, the flow rate of the hydraulic fluid 46 by increasing the lubricant pressure P _J by the lubricating pressure control means 118 is often, the cooling effect of the starting clutch Ci is increased. Alternatively, the lower the engagement rate C _Ci of the starting clutch Ci during engine operation, the higher the temperature of the starting clutch Ci, or the temperatures of the first motor generator MG1 and / or the second motor generator MG2 (T _MG1 , T _MG2). ) May increase and the oil temperature T _OUT of the hydraulic oil 46 may increase. However, since the lubricating pressure control means 118 increases the lubricating pressure P _J to increase the flow rate of the hydraulic oil 46, the hydraulic oil 46 The cooling effect of 46 is enhanced.

Further, according to this embodiment, the lubricant pressure P _J, the so determined based on the temperature state of the first motor generator MG1 and second motor generator MG2, the first motor generator MG1 and / or second motor The flow rate required for cooling the generator MG2 is more appropriately obtained, and the durability of the first motor generator MG1 and / or the second motor generator MG2 is further improved.

For example, the higher the temperature (T _MG1 , T _MG2 ) of the first motor generator MG1 or the second motor generator MG2, the higher the necessity of cooling the first motor generator MG1 or the second motor generator MG2 itself. if increasing the flow rate of the hydraulic fluid 46 by increasing the P _J, cooling effect is further improved. Alternatively, the temperature change of the first motor generator MG1 and second motor generator _{MG2 (T MG1 ', T MG2} ') the greater, if the temperature of the first motor generator MG1 and second motor generator MG2 _{(T MG1, T MG2} ) lubrication pressure P _J even if not as high as it becomes necessary to increase the its temperature (T _{MG1, T MG2)} first becomes relatively high possibility that increases the motor-generator MG1 or the second since the need to cool the motor generator MG2 itself is high, if many the flow rate of the hydraulic fluid 46 by increasing the lubricant pressure P _J, cooling effect is further improved.

Further, according to this embodiment, the lubricant pressure P _J, the so is determined based on the heat generation state of the first motor generator MG1 and second motor generator MG2, the first motor generator MG1 and / or second motor The flow rate required for cooling the generator MG2 is more appropriately obtained, and the durability of the first motor generator MG1 and / or the second motor generator MG2 is further improved.

For example, the higher the heat generation amount (Q _MG1 , Q _MG2 ) of the first motor generator MG1 or the second motor generator MG2 or the greater the change in the heat generation amount (Q _MG1 ′, Q _MG2 ′), Even if the temperature (T _MG1 , T _MG2 ) of one motor generator MG1 or the second motor generator MG2 is not so high that the lubrication pressure P _J must be increased, the temperature (T _MG1 , T _MG2 ) is high. since the need for cooling possibly the first motor generator MG1 and second motor generator MG2 itself relatively high comprised increases, if increasing the flow rate of the hydraulic fluid 46 by increasing the lubricant pressure P _J, cooled The effect is further enhanced.

Further, according to this embodiment, the lubricant pressure P _J, so is determined based on the temperature condition of the hydraulic oil 46, the flow rate required for cooling the first motor-generator MG1 and / or second motor generator MG2 It is obtained more appropriately, and the durability of the first motor generator MG1 and / or the second motor generator MG2 is further improved.

For example, the higher the oil temperature _TOIL of the hydraulic oil 46, the higher the temperature ( _TMG1 , _TMG2 ) of the first motor generator MG1 or the second motor generator MG2 itself, or the first motor generator MG1 or the second motor generator. Since the temperature of MG2 (T _MG1 , T _MG2 ) may be increased, the cooling effect can be further enhanced by increasing the lubricating pressure P _J and increasing the flow rate of the hydraulic oil 46. Alternatively, even if the oil temperature change T _OIL ′ of the hydraulic oil 46 is larger, even if the oil temperature T _OIL of the hydraulic oil 46 is not so high that the lubricating pressure P _J needs to be increased, the oil temperature T _OIL since _{potentially oIL} increases it is relatively high, if many the flow rate of the hydraulic fluid 46 by increasing the lubricant pressure P _J, cooling effect is further improved.

  As mentioned above, although the Example of this invention was described in detail based on drawing, this invention is applied also in another aspect.

For example, in the above-described embodiment, the lubricating pressure determining unit 130 determines the lubricating pressure P _J based on the engagement rate C _Ci of the starting clutch Ci during slip engagement, but in addition to that, the duration of slip engagement is determined. the may be determined lubricating pressure P _J in consideration of. For example, as the its duration is prolonged, there is a high possibility that the temperature rises in the heating value is increased starting clutch Ci and hydraulic oil 46 is relatively high determine lubrication pressure P _J.

In the illustrated embodiment, as shown in FIG. 5, although the hydraulic oil 46 discharged from the line pressure regulating valve 40 to the oil cooler 44 is supplied by the linear solenoid valve SLJ the lubricating pressure P _J pressure regulating The applied hydraulic oil 46 may be supplied directly or via the drive device 6. Thus, lubrication because cooler flow with changes in the pressure P _J also changes, for example when it is likely the absolute value comprised high or when the oil temperature T _OIL high oil temperature T _OIL, the cooler The flow rate is also increased, and the cooling performance of the hydraulic oil 46 itself is further improved.

Further, in the illustrated embodiment, although the hydraulic oil 46 discharged from the line pressure regulating valve 40 is pressure regulated directly lubricated pressure P _J by the linear solenoid valve SLJ, discharged from the example line pressure regulating valve 40 the pressure regulating valve of the relief valve form control pressure is provided to be supplied to the hydraulic oil 46 is controlled from the linear solenoid valve to the lubrication pressure P _J may be pressure regulated to a lubricating pressure P _J.

Further, in the above-described embodiment, the clutch engagement rate correction coefficient calculating unit 126 or the clutch slip rate calculating unit 128 is based on the hydraulic pressure command value from the hybrid control unit 114 and the engagement rate C _Ci or the slip rate C of the start clutch Ci. _{Although SP} is calculated, an actual oil pressure value may be used instead of the oil pressure command value. The rotational speed ratio, or alternatively the first motor-generator rotational speed of the engagement ratio C _Ci and slip ratio C _SP, the relative rotational speed ratio of the starting clutch Ci itself, for example, the crankshaft 9 and the input shaft 16 of the starting clutch Ci N _MG1 and may be calculated based on the ratio of the second motor-generator rotational speed _{N MG2.}

  In the driving device 6 of the above-described embodiment, the crankshaft 9 and the input shaft 16 are exclusively connected via the starting clutch Ci. However, in addition to the starting clutch Ci, a fluid transmission device is provided, The output of the engine 8 may be transmitted to the drive wheels 32 via a transmission device. As the fluid transmission device, for example, a torque converter with a lock-up clutch, a torque converter without a lock-up clutch, or a fluid coupling (fluid coupling) having no torque amplifying action is used.

  In the above-described embodiment, the second motor generator MG2 is connected to the input shaft 16, but may be connected to the output shaft 28 or to a rotating member in the automatic transmission 10.

  Further, the hydraulic friction engagement devices such as the starting clutch Ci, the clutches C1 to C4, the brakes B1 and B2 of the above-described embodiment are a magnetic powder type, an electromagnetic type such as a powder (magnetic powder) clutch, an electromagnetic clutch, and a meshing type dog clutch. It may be composed of a mechanical engagement device.

  The above description is only an embodiment, and the present invention can be implemented in variously modified and improved forms based on the knowledge of those skilled in the art.

BRIEF DESCRIPTION OF THE DRAWINGS FIG. 1 is a skeleton diagram illustrating a main configuration of a vehicle drive device to which a control device according to an embodiment of the present invention is applied. FIG. 2 is an alignment chart for explaining the operation of the automatic transmission of FIG. 1. FIG. 2 is an operation table showing a relationship between a shift stage of the automatic transmission of FIG. 1 and a combination of operations of a hydraulic friction engagement device necessary for establishing the shift stage. FIG. 2 is a diagram illustrating a schematic configuration of the drive device of FIG. 1 and a block diagram illustrating a main part of a control system provided in the vehicle for controlling the drive device and the like. FIG. 5 is a diagram schematically illustrating a configuration of a hydraulic control circuit illustrated in FIG. 4. It is a figure which shows the shift operation apparatus provided in the vehicle of FIG. It is a functional block diagram explaining the principal part of the control function with which the electronic control apparatus of FIG. 4 is provided. FIG. 5 is a diagram showing a relationship between a throttle valve opening and an accelerator operation amount used in electronic throttle valve opening control of the electronic control device of FIG. 4. It is a figure which shows the shift diagram used in the shift control of the electronic controller of FIG. It is a figure explaining an example of the operation mode possible with the vehicle drive device of FIG. FIG. 5 is an example of a relationship (map) that is experimentally determined and determined in advance between an engagement rate of a starting clutch and a correction coefficient when an engine is operated when a vehicle is normally started. It is an example of a relationship (map) that is determined experimentally in advance between the first motor generator temperature or the second motor generator temperature and the lubricating pressure. FIG. 5 is an example of a relationship (map) determined by experimental determination in advance between the first motor generator heat generation amount or the second motor generator heat generation amount and the lubrication pressure. It is an example of a relationship (map) determined by experimental determination in advance between the first motor generator temperature change or the second motor generator temperature change and the lubricating pressure. FIG. 5 is an example of a relationship (map) that is experimentally obtained and determined in advance between a change in the first motor generator heat generation amount or a change in the second motor generator heat generation amount and the lubrication pressure. It is an example of the relationship (map) previously calculated | required experimentally and calculated | required by the oil temperature of the hydraulic oil at the time of vehicle starting normally implemented, or oil temperature change, and a correction coefficient. FIG. 5 is an example of a relationship (map) determined experimentally in advance between the oil temperature of hydraulic oil at the time of friction start or a change in oil temperature and the lubricating pressure. It is an example of a relationship (map) that has been experimentally determined and determined in advance between the slip ratio of the starting clutch and the correction coefficient at the time of friction start. The main part of the control operation of the electronic control unit of FIG. 4, that is, the lubricating pressure of the hydraulic oil circulated and supplied to the drive device at the start of the vehicle is controlled so that the flow rate of the hydraulic oil necessary for cooling can be obtained appropriately. It is a flowchart explaining a control action.

Explanation of symbols

6: Drive device (power transmission device)
8: Engine 90: Electronic control device (lubrication control device)
98: Hydraulic control circuit (fluid circulation supply device)
118: Lubrication pressure control means (lubrication flow rate control means)
Ci: Starting clutch MG1: First motor MG2: Second motor

Claims (8)

A first motor that functions as a generator by being rotationally driven by an engine, a second motor that is rotationally driven by the generated energy of the first motor and functions as a driving force source for traveling, the engine, and the first motor And a starting clutch that connects the second electric motor to be connectable / disconnectable, and a lubricating fluid is circulated and supplied to the power transmitting device, and the first electric motor and the second electric motor are driven by the lubricating fluid. A fluid circulation supply device for cooling, a vehicle lubrication control device comprising:
Lubricating pressure control means for controlling the lubricating pressure of the lubricating fluid circulated and supplied to the power transmission device by the fluid circulation supply device based on the torque capacity of the starting clutch when the engine and the second electric motor are operating, seen including,
The lubricating pressure control means corrects the lubricating pressure of the lubricating fluid to be smaller as the driving force based on the output of the engine transmitted to the driving wheels is relatively increased according to the torque capacity of the starting clutch. A vehicle lubrication control device.   The vehicle lubrication control device according to claim 1, wherein the lubrication pressure is determined based on a temperature state of the first electric motor or the second electric motor.   3. The vehicle lubrication control device according to claim 1, wherein the lubrication pressure is determined based on a heat generation state of the first electric motor or the second electric motor. The vehicle lubrication control device according to any one of claims 1 to 3, wherein the lubricating pressure is determined based on a temperature state of the lubricating fluid. A first motor that functions as a generator by being rotationally driven by an engine, a second motor that is rotationally driven by the generated energy of the first motor and functions as a driving force source for traveling, the engine, and the first motor And a starting clutch that connects the second electric motor to be connectable / disconnectable, and a lubricating fluid is circulated and supplied to the power transmitting device, and the first electric motor and the second electric motor are driven by the lubricating fluid. A fluid circulation supply device for cooling, a vehicle lubrication control device comprising:
Lubrication flow rate control means for controlling the lubrication flow rate of the lubricating fluid circulated and supplied to the power transmission device by the fluid circulation supply device based on the torque capacity of the starting clutch when the engine and the second electric motor are operating, seen including,
The lubrication flow rate control means corrects the lubrication flow rate of the lubricating fluid to be smaller as the driving force based on the output of the engine transmitted to the drive wheels according to the torque capacity of the starting clutch is relatively increased. A vehicle lubrication control device.   The vehicle lubrication control device according to claim 5, wherein the lubrication flow rate is determined based on a temperature state of the first electric motor or the second electric motor.   The vehicle lubrication control device according to claim 5 or 6, wherein the lubrication flow rate is determined based on a heat generation state of the first electric motor or the second electric motor. The vehicle lubrication control device according to any one of claims 5 to 7, wherein the lubrication flow rate is determined based on a temperature state of the lubricating fluid.

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