Torque overlay device for a hybrid drive system, and a method for operating such a hybrid drive system

Abstract

The invention relates to a torque overlay device for use in a hybrid drive system for motor vehicles having an internal combustion engine, an electric motor and the torque overlay device, wherein a torque of the internal combustion engine and of the electric motor are overlaid using the torque overlay device. The torque overlay device is connected, on the output side, to a driven element of the vehicle, and includes a first and a second torque input and a torque output, as well as a first transmission device and a second transmission device. The electric motor can be coupled to the first torque input in a torque-resistant manner, and the internal combustion engine can be coupled to the second torque input in a torque resistant manner, where the first torque input is connected to the first transmission device and the second torque input is connected to the second transmission device, both in a torque-resistant manner. In addition, the transmission devices may each be coupled, on the driven side, to the torque output of the torque overlay device in a torque-resistant manner.

Claims

1. A torque superimposition device of a hybrid drive of a vehicle, the torque superimposition device comprising: a first torque input torque-proof coupled to an electric motor of the hybrid drive; a second torque input torque-proof coupled to an internal combustion engine of the hybrid drive; a torque output coupled to a drive input of the vehicle; a first gearing device torque-proof coupled to the first torque input and torque-proof coupled to the torque output, the first gearing device comprising: at least two alternatively selectable gear ratio steps, at least one planetary gear set configured to implement the at least two alternatively selectable gear ratio steps; and a second gearing device torque-proof coupled to the second torque input and torque-proof coupled to the torque output, the second gearing device comprising: at least one alternatively selectable gear ratio step, and at least one spur gear set configured to implement the at least one selectable gear ratio step, wherein torque of the electric motor is superimposed on torque of the internal combustion engine via the torque superimposition device.

2. The torque superimposition device according to claim 1, wherein each of the first and second gearing devices has at least two alternatively selectable gear ratio steps.

3. The torque superimposition device according to claim 1, wherein the second gearing device comprises at least two spur gear sets implementing the at least two alternatively selectable gear ratio steps.

4. The torque superimposition device according to claim 1, wherein the first gearing device comprises exactly one planetary gear set, wherein exactly two gear ratio steps can be alternatively selected.

5. The torque superimposition device according to claim 1, wherein the planetary gear set is configured with at least one of a power-shift friction brake and a friction clutch to select the gear ratio steps.

6. The torque superimposition device according to claim 1, wherein the planetary gear set is configured as a block to implement one of the gear ratio steps of the at least two alternatively selectable gear ratio steps.

7. The torque superimposition device according to claim 1, wherein the electric machine is coupled to a ring gear of the planetary gear set in a torque-proof manner, a sun gear of the planetary gear set is coupled to a housing of the torque superimposition device in a stationary manner by the power-shift friction brake, and a planet carrier of the planetary gear set is coupled to the torque output in a torque-proof manner in order to implement a further gear ratio step of the at least two alternatively selectable gear ratio steps.

8. The torque superimposition device according to either claim 7, wherein at least one of a form-locked and frictionally engaged clutch is arranged on the shaft, and the clutch is implemented with a synchronizer element so as to alternatively selectably connect the shaft to the fixed gearing element in a torque-proof manner.

9. The torque superimposition device according to claim 1, wherein the second gearing device comprises at least one form-locked and frictionally engaged clutch configured for the selection of the at least one spur gear set configured as a gear ratio step.

10. The torque superimposition device according to claim 9, wherein the at least one clutch is implemented with a synchronizer element.

11. The torque superimposition device according to claim 1, wherein a fixed gearing element is arranged between the electric motor and the planetary gear set, a spur gear set having a first spur gear facing the electric motor and a second spur gear facing the planetary gear set, wherein the fixed gearing element has a gear ratio between 1.5 and 3.

12. The torque superimposition device according to claim 11, wherein a shaft of the second gearing device, which is connected to the second torque input, is connectable to the first spur gear of the fixed gearing element configured as the spur gear set in an alternatively selectable and torque-proof manner so as to implement one or two further alternatively selectable gear ratio steps.

13. The torque superimposition device according to claim 1, wherein the shaft of the second gearing device, which is connected to the second torque input, and a shaft of the first gearing device, which is connected to the first torque input, are arranged aligned with each other, and wherein the end of the shaft of the second gearing device, which faces away from the second torque input, is mounted in an end of the shaft of the first gearing device, which is designed as the hollow shaft, said shaft being connected to the first torque input.

14. A hybrid drive of a vehicle, the hybrid drive comprising: an internal combustion engine having a power between 40 and 150 kW, an electric motor having a continuous power of between 30 and 60 kW, and a peak power of 2 to 3 times the continuous power, and a torque superimposition device comprising: a first torque input torque-proof coupled to the electric motor; a second torque input torque-proof coupled to the internal combustion engine; a torque output coupled to a drive input of the vehicle; a first gearing device torque-proof coupled to the first torque input and torque-proof coupled to the torque output, the first gearing device comprising: at least two alternatively selectable gear ratio steps, and at least one planetary gear set configured to implement the at least two alternatively selectable gear ratio steps; and a second gearing device torque-proof coupled to the second torque input and torque-proof coupled to the torque output, the second gearing device comprising: at least one alternatively selectable gear ratio step, and at least one spur gear set configured to implement the at least one selectable gear ratio step, wherein torque of the electric motor is superimposed on torque of the internal combustion engine via the torque superimposition device.

Description

BRIEF DESCRIPTION OF THE DRAWINGS

(1) FIG. 1 shows an embodiment of the torque superimposition device according to the invention;

(2) FIG. 2 shows a normal hybrid operating mode;

(3) FIG. 3 shows a first special hybrid operating mode;

(4) FIG. 4 shows an all-electric driving mode; and

(5) FIG. 5 shows an internal combustion engine-based driving mode.

DETAILED DESCRIPTION OF THE DRAWINGS

(6) Only elements and components that are essential for gaining an understanding of the invention are shown in the figures. The shown exemplary embodiment shall be understood to have a purely instructive nature and is intended to provide a better understanding, without limiting the subject matter of the invention.

(7) FIG. 1 shows a hybrid drive 1 implemented according to the invention for use in a passenger motor vehicle. The hybrid drive 1 comprises an internal combustion engine VM, an electric motor EM, and a torque superimposition device implemented according to the invention as a transmission 2. The transmission 2 comprises a first torque input 3 and a second torque input 4, as well as a torque output 5.

(8) The electric motor EM is coupled to the first torque input 3 in a torque-proof manner, wherein the torque-proof coupling is achieved by means of a first shaft 6 here. For this purpose, the first shaft 6 is coupled to the rotor shaft (not shown in FIG. 1) of the electric motor EM in a torque-proof manner. Corresponding couplings and/or coupling elements for coupling two shafts are known in the prior art.

(9) The internal combustion engine VM is coupled to the second torque input 4 in a torque-proof manner, wherein the torque-proof coupling is achieved by means of a second shaft 7 here. For this purpose, the second shaft 7 is coupled to the crankshaft (not shown in FIG. 1) of the internal combustion engine VM in a torque-proof manner. Corresponding couplings and/or coupling elements for coupling two shafts are known in the prior art.

(10) The torque output 5 of the transmission 2 is coupled to an input spur gear 8 of a spur gear differential D of the output 9 of the vehicle in a torque-proof manner, wherein the coupling is achieved by means of an interference fit of the input spur gear 8 here. The spur gear differential D is implemented with a gear ratio between 3 and 4 here. The output 9 furthermore comprises the output shaft 10 and the wheels 11a and 11b. The hybrid drive according to the invention can be used as a front-wheel drive or, in principle a rear-wheel drive, standard drive, or even all-wheel drive. A front-wheel drive is shown in FIG. 1.

(11) The transmission 2 implemented according to the invention comprises a first gearing device, which is implemented as a first subtransmission 12 here, and a second gearing device, which is implemented as a second subtransmission 13 here.

(12) The first shaft 6 is implemented as part of the first subtransmission 12 here, and the second shaft 7 is implemented as part of the second subtransmission 13. In this way, the two subtransmissions are coupled to the respective propulsion systems EM and VM in a torque-proof manner via the respective torque inputs 3 and 4.

(13) The first subtransmission 12 moreover comprises a fixed gearing element, which is designed as a spur gear set 14 here, as well as exactly one planetary gear set 15.

(14) The spur gear set 14 is designed as a fixed gearing element and comprises a first spur gear 16 facing the electric motor EM and a second spur gear 17 facing the planetary gear set 15.

(15) The first spur gear 16 of the spur gear set 14, which is designed as a fixed gearing element, is arranged at the end of the shaft 6 facing away from the torque input 3 on this shaft 6 by means of interference fit. The end of the shaft 6 facing away from the torque input 3 is implemented as a hollow shaft 18 in an end section of the shaft 3 here.

(16) The second spur gear 17 of the spur gear set 14 designed as a fixed gearing element is arranged on a cylindrical outlet 19 of the ring gear 20 of the planetary gear set 15 by means of interference fit.

(17) The spur gear set 14 designed as a fixed gearing element is implemented with a gear ratio of 2 here.

(18) The exactly one planetary gear set 15 comprises the ring gear 20, the planet gears 21, the planet carrier 22, and the sun gear 23. The sun gear 23 can be braked with respect to a housing (not shown in FIG. 1) of the transmission 2 by means of the power-shift friction brake 24 in a stationary manner.

(19) If braking of the sun gear 23 takes place in a stationary manner by means of the friction brake 24, the ring gear 20 and the planet gears 21, together with the coupled planet carrier 22, run on the stationary sun gear 23, whereby a torque that is introduced by the electric motor EM via the fixed gearing element 14 with a rotational speed is assigned a speed ratio and torque ratio corresponding to the gear ratio between the ring gear 20 and planet gears 21 or planet carrier 22, and is transmitted to the output shaft 25 via the planet carrier 22. By braking the sun gear 23, a first gear EM1 can therefore be implemented as a gear ratio step of the first subtransmission 12. The gear ratio of the planetary gear set when the sun gear 23 is braked in a stationary manner is between 1.5 and 1.9 here, preferably 1.8.

(20) The second gear EM2, serving as the second gear ratio step of the first subtransmission 12, can be implemented by a block configuration of the planetary gear set 15. For this purpose, the clutch 26 between the ring gear 20 and planet carrier 22 is closed, whereby the entire planetary gear set 15 is configured as a block and rotates as a whole. The gear ratio of the planetary gear set 15 is therefore 1. A torque that is introduced by the electric motor EM via the fixed gearing element 14 with a rotational speed is assigned a speed ratio and torque ratio using the gear ratio of 1, and is transmitted to the output shaft 25 via the planet carrier 22.

(21) In the exemplary embodiment shown in FIG. 1, exactly two gears can therefore be implemented as two alternatively selectable gear ratio steps of the first subtransmission 12.

(22) The shifting transitions from first gear to second gear of the first subtransmission 12, and conversely, can be achieved by closing or opening the clutch 26, while simultaneously opening or braking the friction brake 24 under load, and can therefore be implemented without any drop in tractive force.

(23) If both the clutch 26 as well as the friction brake 24 are open, no coupling of moment to the output shaft 25 exists. This operating state is of interest for the “stationary charging” function when the vehicle is stationary, with power flowing from the internal combustion engine VM to the electric motor EM operated as a generator in this operating state.

(24) The output shaft 25 is designed as a torque output 5 of the transmission 2 on the side facing away from the planetary gear set 15.

(25) The second subtransmission 13 moreover comprises two spur gear sets 27 and 28 so as to implement a first and a second alternatively selectable gear ratio step. The spur gear sets are implemented as an idler gear/fixed gear combination known from the prior art, wherein the idler gears of the spur gear sets 27 and 28 arranged on the second shaft 7 can be alternatively connected by means of a dog clutch 29 that acts on both sides and can be axially displaced on the shaft 7. The dog clutch 29 is composed of two sub dog clutches, which are both implemented as form-locked clutches. The two sub dog clutches moreover in each case additionally comprise a synchronizer element, which is a single-cone synchronizer here (not shown in FIG. 1).

(26) As a result of form-locked engagement of the dog clutch 29 in the idler gear of the spur gear set 27, a torque-proof positive fit is established between the second shaft 7 and the spur gear set 27. Corresponding to the gear ratio of the spur gear set 27, a torque that is introduced by the internal combustion engine VM with a rotational speed is assigned a speed ratio and torque ratio and is transmitted via the fixed gear of the spur gear set 27 to the output shaft 25 rigidly connected to the fixed gear, and from there to the spur gear differential D.

(27) The same mechanism of action applies to the spur gear set 28.

(28) The gear ratio of the spur gear set 27 is between 1 and 1.5 for the exemplary embodiment shown here, and that of the spur gear set 28 is between 0.6 and 1.2. The gear ratio step comprising the spur gear set 27 is designed as the third gear VM3 of the second subtransmission 13 here, and the gear ratio step comprising the spur gear set 28 is designed as the fourth gear VM4 of the second subtransmission 13.

(29) To implement the first gear VM1 and the second gear VM2 as two further alternatively selectable gear ratio steps of the second subtransmission, the second shaft 7 can be connected in a torque-proof manner to the first spur gear 16 of the spur gear set 14 designed as a fixed gearing element by means of form-locked dog clutches 30, alternatively to gears 3 and 4. If the dog clutch 30 is connected in a form-locked manner to the spur gear set 14, a torque supplied by the internal combustion engine VM is transmitted from the second shaft 7 via the spur gear set 14 to the planetary gear set 15. Depending on the shift state of the planetary gear set 15, the torque is transmitted corresponding to the gear ratio VM1=EM1 or VM2=EM2 to the output shaft 25, and from there to the spur gear differential D and the output 9.

(30) The dog clutch 30 is configured with a single-cone synchronizer here.

(31) VM1 is implemented as a short starting gear here. A starting clutch arranged directly downstream of the internal combustion engine VM is not shown in FIG. 1. A person skilled in the art will be familiar with the design and arrangement of such a starting clutch. VM2 is implemented as an intermediate gear on the driving power side here for the implementation of high-load or full-load maneuvers with aggregate power of the propulsion systems. VM3 is designed as a vmax gear here, and VM4 as an energy efficiency gear.

(32) The gear ratios of the individual gear ratio steps are preferably in the following ranges: gear ratio of differential 9: approximately 2.5 to 4, particularly preferably 3.5 to 4 gear ratio of planetary gear set 15: block configuration: 1 with sun gear 23 braked in a stationary manner with respect to housing 24: approximately 1.5 to 2, particularly preferably 1.7 to 1.8 gear ratio of fixed gearing element 14: approximately 1.5 to 3, particularly preferably 1.5 to 2 gear ratio of spur gear set 27: approximately 0.9 to 1.4, particularly preferably 1 to 1.2. gear ratio of spur gear set 28: approximately 0.6 to 0.8, particularly preferably 0.7

(33) The following preferred gear ratio ranges apply to the gears: gear ratio of EM1: approximately 11 to 14, particularly preferably approximately 12 gear ratio of EM2: approximately 6 to 9, particularly preferably 6 to 7 gear ratio of VM1: like gear ratio of EM1 (approximately 11 to 14, particularly preferably approximately 12) gear ratio of VM2: like gear ratio of EM2 (approximately 6 to 9, particularly preferably 7 to 8) gear ratio of VM3: approximately 3.5 to 5, particularly preferably 3.5 to 4 gear ratio of VM4: approximately 2.5 to 3.5, particularly preferably approximately 3

(34) According to FIG. 1, the second shaft 7 and the first shaft 6 are arranged aligned with each other here, wherein the end of the second shaft 7 facing away from the internal combustion engine VM is mounted in the end of the first shaft 6 designed as the hollow shaft 18.

(35) The hybrid drive implemented in FIG. 1 has a very compact design. With respect to gears VM3 and VM4, as well as EM1 and EM2, the propulsion systems can be individually activated in a propulsion system-specific manner, whereby high driving dynamics and high energy efficiency are achieved. Due to the very compact design of the transmission having few gear meshing events in the respective moment paths, an extremely good efficiency of the transmission is achieved compared to transmissions known from the prior art. Moreover, the shifting processes between EM1 and EM2 (and also between VM1 and VM2) can be implemented without interruption of torque flow. Shifting processes from VM2 to VM3 and VM4 are carried out with assistance from the electric machine EM, and thus are carried out with sufficiently good shifting quality.

(36) The hybrid drive of a motor vehicle, which comprises the above-described torque superimposition device, is advantageously operated so as to comprise a normal hybrid operating mode having a first and a second control range, as shown in FIG. 2.

(37) The first control range covers a vehicle speed of zero up to an electric driving speed limit, for example approximately 60 km/h. The second control range covers the electric driving speed limit up to the maximum speed of the vehicle.

(38) It is advantageous to define the electric driving speed limit as a connecting and disconnecting threshold for the internal combustion engine. The connecting threshold is then the driving speed value at which the switch is made from the first control range to the second control range as the driving speed increases. The disconnecting threshold is then the driving speed value at which the switch is made from the second control range to the first control range as the driving speed decreases.

(39) The connecting threshold and disconnecting threshold are usually approximately 10 to 15 km/h apart from each other.

(40) In the first control range, the drive torque of the motor vehicle is generated only by means of the electric motor EM, and in the second control range, the drive torque of the motor vehicle is generated in hybrid operation by the cooperation of the internal combustion engine VM and the electric motor EM.

(41) The division into two control ranges makes it possible to optimally design the drive configuration in each case from an energy point of view. From an energy efficiency view, it is therefore advantageous to drive by means of the electric motor in the first control range. Here, the electric motor has a higher efficiency than would be possible with the internal combustion engine propulsion system. If additionally the amount of energy required for the electric motor-based operation was made available via charging from a power grid, this can result in ecological CO2 neutrality of these driving states. In addition to energy efficiency, the emission-free operation, such as in inner-city areas, is another criterion to select the electric driving speed limits. With respect to both criteria, an electric driving speed in the range of approximately 60 km/h constitutes a good compromise during normal operation.

(42) In the second control range, on the other hand, it is more advantageous from an energy efficiency and efficiency point of view to drive in a combined operating mode, this being the interaction between internal combustion engine and electric motor.

(43) It is particularly advantageous from an energy point of view to consistently operate the internal combustion engine in continuous operation, for example along the beopt line of the same, and to supply the brief dynamics required for the driving task via the electric motor. Dynamics here shall be understood to mean both positive dynamics within the meaning of an acceleration of the vehicle, as well as negative dynamics within the meaning of a retardation of the vehicle.

(44) In addition to optimal energy efficiency of the hybrid drive in each operating and driving state of the vehicle, the method according to the invention also ensures that vehicles comprising such a hybrid drive have a total range that is comparable to conventional vehicles operated by internal combustion engines, while also having a high electric range, whereby such a vehicle is suitable for both city operation, where optionally entry restrictions may exist for operation only by means of electric motor, and for long distances, without necessitating complex and time-consuming electricity “refueling.”

(45) The generation of the drive torque in the first control range according to a control strategy according to the invention differs from methods of comparable hybrid drives that are typically known from the prior art also in that no activation of the combustion engine for power compensation is permitted even during acceleration processes. On the one hand, the operation of the first control range according to the invention ensures that driving takes place exclusively by means of the electric motor in the first control range; on the other hand, such activation of the combustion engine for power compensation known from the prior art can result in a worsening of the efficiency and in problems regarding emissions and warm-up behavior of the internal combustion engine.

(46) If the power requirement exceeds the power provision of the electric motor (for example, launch control function), activation of the internal combustion engine for power compensation can take place during acceleration phases within the meaning of a special function for optimal use of the aggregate power so as to achieve maximal driving power.

(47) The internal combustion engine VM is operated independently of the load requirement of the driver input in the range of optimal efficiency of the internal combustion engine VM, or close thereto. The control according to the invention of the internal combustion engine-based operation at optimal efficiency of the internal combustion engine also includes setting the internal combustion engine-based operating point to an optimal system efficiency, which results from the efficiency with internal combustion engine operation at the respective internal combustion engine-based operating point, and a charge/discharge efficiency of the electric energy storage unit at the charge or discharge power of the high-voltage battery system resulting at the respective internal combustion engine-based operating point, the efficiency of the control electronics and of the electric motor when operated as a generator, and the respective driving task.

(48) It is therefore possible that a lower efficiency than the maximum possible efficiency (beopt efficiency) of the combustion engine is set so as to optimize the system efficiency. This is particularly the case when a considerable charging rate would be generated if the combustion engine were set to the maximum possible efficiency, which would result in a disproportionate degradation of the efficiency of the high-voltage battery system.

(49) Moreover, a special driving power operating mode is possible. It is only in this special driving power operating mode that the internal combustion engine VM is operated briefly at loads above the range of optimal efficiency of the internal combustion engine VM. In this way, it becomes possible for the maximum power of the internal combustion engine available in this driving state of the vehicle to be briefly accessed, even if this results in a deterioration of the efficiency of the internal combustion engine-based operation.

(50) The method according to the invention moreover comprises a first special hybrid operating mode, which replaces the normal hybrid operating mode as a function of the battery charge state. This first special hybrid operating mode is activated when the battery charge state drops below a first limit value. Upon activation of the first special hybrid operating mode, the electric driving speed limit is set to a value lower than the original value.

(51) The transition from normal hybrid operating mode to the first special hybrid operating mode can also take place continuously or steplessly. In particular, such continuous or stepless setting of the respective operating mode can also be carried out based on information about the desired driving route and the current traffic situation (anticipatory function).

(52) Overall, this means that a first control range, in which driving takes place solely by means of the electric motor, is reduced toward lower speeds as the battery charge state decreases. In the extreme case of a completely discharged battery state, this could cause the transition into the second control range to already take place at 0 km/h. In addition, the operating point of the combustion engine can be shifted above the range of optimal efficiency toward full load when the battery charge state is low and the charging balance is negative, so as to generate a higher charging rate.

(53) Conventional operating strategies typically degrade in this case via the load, which is to say that a switch into hybrid operation is already made with a lower load requirement when the battery charge state decreases.

(54) The method moreover comprises a second special hybrid operating mode, which replaces the normal hybrid operating mode as a function of the battery charge state. The second special hybrid operating mode is activated when the battery charge state exceeds a second limit value and a charging rate of the internal combustion engine is greater than zero. Upon activation of the second special hybrid operating mode, the electric driving speed limit can be set to a value higher than the original value and/or the operation of the internal combustion engine VM is adjusted to a load below the range of optimal efficiency of the internal combustion engine VM, in such a way that the charging rate by the internal combustion engine VM is equal to zero.

(55) The transition from normal operating mode to the second special operating mode also advantageously takes place continuously or steplessly, and particularly advantageously by including anticipatory functions and/or as a function of the battery charge state.

(56) In addition to the system efficiencies, the battery charge state is a criterion in both operating modes for the selection of the operating point of the internal combustion engine, and therefore also for the charging rate. If battery charge states are very high (optionally in conjunction with a positive average charging balance), the charging rate is reduced to zero or into the negative range; if battery charge states are very low (optionally in conjunction with a negative average charging balance) it is increased up to full load of the combustion engine, even if this results in disadvantages in the efficiency of the combustion engine.

(57) In many cases it will be advantageous for fuel to continue to be supplied to the internal combustion engine at a constant operating point in the second control range even in coasting operation, which is referred to as fired coasting. Fired coasting is in particular of advantage when a large amount of charging energy must be generated over a short time.

(58) Moreover, the method comprises an electric driving operating mode which the driver can optionally select, as shown in FIG. 4, and in which the drive torque of the motor vehicle is generated only by means of the electric motor EM across the entire speed range.

(59) Finally, as shown in FIG. 5, a pure internal combustion engine-based operation can also be implemented. If high shifting comfort is desired, the electric motor has to assume the assistance function during the shifting process only briefly in a pure internal combustion engine-based operation.

(60) The above description of the present invention serves only illustrative purposes and is not intended to restrict the invention. Within the context of the invention, various changes and modifications are possible without departing from the scope of the invention or the equivalents thereof.

LIST OF REFERENCE NUMERALS

(61) 1 hybrid drive 2 transmission 3 first torque input 4 second torque input 5 torque output 6 first shaft 7 second shaft 8 input spur gear 9 output 10 driven shaft 11a wheels 11b wheels 12 first subtransmission 13 second subtransmission 14 spur gear set 15 planetary gear set 16 first spur gear 17 second spur gear 18 hollow shaft 19 outlet 20 ring gear 21 planet gears 22 planet carrier 23 sun gear 24 friction brake 25 output shaft 26 clutch 27 spur gear set 28 spur gear set 29 dog clutch 30 dog clutch D spur gear differential EM electric motor EM1 first gear EM2 second gear VM internal combustion engine VM1 first gear VM2 second gear VM3 third gear VM4 fourth gear

(62) The foregoing disclosure has been set forth merely to illustrate the invention and is not intended to be limiting. Since modifications of the disclosed embodiments incorporating the spirit and substance of the invention may occur to persons skilled in the art, the invention should be construed to include everything within the scope of the appended claims and equivalents thereof.