GB2530396A

GB2530396A - Hydraulic system for a vehicle, in particular a motor vehicle

Info

Publication number: GB2530396A
Application number: GB1513761.5A
Authority: GB
Inventors: Ethan Ott; Michael Norlin; Shivkumar Duraiswamy
Original assignee: Daimler AG
Current assignee: Mercedes Benz Group AG
Priority date: 2015-08-04
Filing date: 2015-08-04
Publication date: 2016-03-23
Also published as: GB201513761D0

Abstract

A hydraulic system 10 for a vehicle, typically a truck or other commercial vehicle, comprises a pump 14 drivable by an engine and decouplable from the engine using a clutch 20. Pump 14 pumps hydraulic fluid through two hydraulic circuits 22, 28, with a valve 32 distributing fluid through circuits 22, 28. First circuit 22 comprises a hydraulic accumulator 24 and a closed centre steering gear 26 of a power steering system. Second circuit 28 comprises a hydraulic motor 30, which may drive a fan 36. Pump 14 may be driven through an air compressor 16 driven by the engine. A cooler 40 may be provided in one of the circuits 22, 28 to cool the hydraulic fluid. Hydraulic system 10 may be operated efficiently, with low parasitic load on the engine when steering assistance is not required.

Description

Hydraulic System for a Vehicle, in Particular a Motor Vehicle The invention relates to a hydraulic system for a vehicle, in particular a commercial vehicle, as well as a method for operating a hydraulic fluid of a vehicle.

US s 100 221 B2 shows a rotary valve for a center closed power steering system, the rotary valve comprising a cylindrical sleeve with an input port and an output port. The valve further comprises a shaft rotatably positioned within the sleeve wherein a land positioned on an outer surface of the shaft is configured to sealingly engage an inner surface of the sleeve in a first, central rotational position of the shaft within the sleeve, thereby preventing flow of fluid between the input port and the output port in the first, central rotational position of the shaft, but not sealingly engage the inner surface of the sleeve in a second rotational position of the shaft within the sleeve, thereby allowing a flow of fluid between the input port and the output port in the second rotational shaft position.

Furthermore, the valve comprises a V-shaped notch on an input port side of the land, said notch providing an opening, for flow of fluid between the input port and the output port as the shaft rotates from the first rotational position to the second rotational position, with said opening beginning at a near-zero width opening at the beginning point of the V and increasing in width along the longitudinal axis of the valve as the shaft rotates from the first rotational position to the second rotational position.

Conventionally, power train cooling system fan drives and hydraulic and pneumatic pumping systems mechanically coupled to an engine of a vehicle consume more energy parasitically generating heat then supplying usable mechanical power if they were sized for extreme situations and then predominantly operated well below their peak efficiency range.

It is therefore an object of the present invention to provide a hydraulic system for a vehicle as well as a method for operating a hydraulic system of a vehicle, by means of which hydraulic system and method the vehicle power consumption efficiency can be improved.

This object is solved by a hydraulic system having the features of patent claim 1 as well as a method having the features of patent claim 9. Advantageous embodiments with expedient developments of the invention are indicated in the other patent claims.

The invention relates to a hydraulic system for a vehicle, in particular a motor vehicle such as, for example, a commercial vehicle. The hydraulic system comprises a pump configured to pump a hydraulic fluid. Moreover, the pump is configured to be driven by an engine of the vehicle. For example, said engine is configured to drive the vehicle.

The hydraulic system further comprises a clutch configured to couple the pump to and decouple the pump from the engine. When the pump is coupled to the engine by means of the clutch the pump can be driven by the engine which is, for example, configured as an internal combustion engine. When the pump is decoupled from the engine by means of the clutch the pump cannot be driven or is not driven by the engine even when the engine is running.

The hydraulic system according to the present invention further comprises a first hydraulic circuit through which a hydraulic fluid can flow. At least one hydraulic accumulator of the hydraulic system is arranged in the first hydraulic circuit, the hydraulic accumulator being configured to store the hydraulic fluid. Particularly, the hydraulic accumulator is configured to store the hydraulic fluid under pressure. The hydraulic system further comprises a closed center steering gear of a power steering system, the closed center steering gear being arranged in the first hydraulic circuit so that the closed center steering gear can be supplied with the hydraulic fluid flowing through the first hydraulic circuit. Thus, a hydraulic power steering system can be realized by means of the closed center steering gear and the hydraulic fluid flowing through the first hydraulic circuit.

The hydraulic system according to the present invention further comprises a second hydraulic circuit through which the hydraulic fluid can flow. At least one hydraulic motor is arranged in the second hydraulic circuit, the hydraulic motor being configured to be driven by the hydraulic fluid flowing through the second hydraulic circuit. Moreover, the hydraulic system according to the present invention comprises a valve device configured to distribute the hydraulic fluid to the first and second hydraulic circuits. In other words, the hydraulic fluid pumped by the pump can be distributed or supplied to the first and second hydraulic circuits by means of the valve device. The hydraulic system according to the present invention allows realizing a particularly high power consumption efficiency of the vehicle since both the first hydraulic circuit and the second hydraulic circuit and, thus, the closed center steering gear and the motor can be supplied with and driven by the hydraulic fluid pumped by the pump. In other words, the hydraulic motor and the closed center steering gear can be supplied with hydraulic fluid by means of the same pump which can be driven by the engine in a need-based manner since the pump can be coupled to and decoupled from the engine by means of the clutch.

For example, the hydraulic system comprises an electronic control unit configured to control the clutch so as to realize a computer controlled clutching system to minimize low and zero efficiency operation of the pump. By means of the hydraulic accumulator an uninterrupted hydraulic power supply to the closed center steering gear can be realized even if the pump is decoupled from the engine or the engine is switched off.

The invention further relates to a method for operating a hydraulic system of a vehicle. In said method a hydraulic fluid pumped by at least one pump of the hydraulic system is distributed by means of a valve device, wherein the hydraulic fluid is distributed to a first hydraulic circuit and a second hydraulic circuit of the hydraulic system. The first hydraulic circuit comprises at least one hydraulic accumulator configured to store the hydraulic fluid flowing through the first hydraulic circuit. This means the hydraulic accumulator is arranged in the first hydraulic circuit. Moreover, the first hydraulic circuit comprises a closed center steering gear of a power steering system so that the closed center steering can be supplied with hydraulic fluid flowing through the first hydraulic circuit.

Furthermore, the second hydraulic circuit comprises at least one hydraulic motor being configured to be driven by the hydraulic fluid flowing through the second hydraulic circuit.

Hence, by means of the valve device, the hydraulic circuits and, thus, the hydraulic accumulator, the closed center steering gear, and the hydraulic motor can be supplied with the hydraulic fluid in an efficient and need-based manner. Advantages and advantageous embodiments of the hydraulic system according to the present invention are to be regarded as advantages and advantageous embodiments of the method according to the present invention and vice versa.

Further advantages, features, and details of the invention derive from the following description of a preferred embodiment as well as from the drawings. The features and feature combinations previously mentioned in the description as well as the features and feature combinations mentioned in the following description of the figures and/or shown in the figures alone can be employed not only in the respectively indicated combination but also in other combination or taken alone without leaving the scope of the invention.

The drawings show in: Fig. 1 a schematic representation of a hydraulic system according to the present invention; Fig. 2 a diagram used to illustrate the functionality of a hydraulic accumulator of the hydraulic system; Fig. 3 a diagram used to illustrate the functionality of a hydraulic power steering system of the hydraulic system; Fig. 4 two diagrams used to illustrate the functionality of the hydraulic system; Fig. 5 a flow diagram illustrating a functionality of the hydraulic system; Fig. 6 afurtherflow diagram illustrating a functionality of the hydraulic system; Fig. 7 afurtherflow diagram illustrating a functionality of the hydraulic system; Fig. 8 a further diagram illustrating a functionality of the hydraulic system; Fig. 9 afurtherflow diagram illustrating a functionality of the hydraulic system; and Fig. 10 a further diagram illustrating a functionality of the hydraulic system.

In the figures the same elements or elements having the same functions are indicated by the same reference signs.

Fig. 1 shows a schematic representation of a hydraulic system 10 of a vehicle in the form of a motor vehicle. For example, the motor vehicle is a commercial vehicle such as a truck. The motor vehicle comprises an engine which is not shown in Fig. 1. For example, the engine is configured as an internal combustion engine, wherein the engine is configured to drive the vehicle. In Fig. 1, a portion of a drive shaft 12 of the engine is shown, wherein the engine provides power or torques via the drive shaft 12.

The hydraulic system 10 comprises a pump configured as a variable displacement pump 14. Moreover, the hydraulic system 10 comprises at least one air compressor 16 configured to compress air for a compressed air system of the vehicle. The air compressor 16 comprises at least one cylinder and at least one piston which is arranged in the cylinder. The piston is translationally movable in the cylinder and pivotably coupled to a crankshaft 18 of the air compressor 16. As can be seen from Fig. 1. the crankshaft iBis configured as a through shaft, wherein the variable displacement pump 14 is coupled to the crankshaft 18. Moreover, the hydraulic system 10 comprises a clutch 20 configured to couple the crankshaft 18 and, thus, the air compressor 16 to and decouple the crankshaft 18 and, thus, the air compressor 16 from the drive shaft 12 and, thus, the engine. In other words, the crankshaft 18 and, thus, the air compressor 16 can be driven by the engine when the clutch 20 is closed. When the clutch 20 is open, the crankshaft 18 and, thus, the air compressor 16 cannot be driven by the engine. Since the variable displacement pump 14 is coupled to the crankshaft 18 and, thus, the air compressor 16, the clutch 20 is configured to couple the variable displacement pump 14 to and decouple the variable displacement pump 14 from the engine by closing and opening the clutch 20.

This means the variable displacement pump 14 can be driven by the engine via the crankshaft 16 and, thus, the air compressor 16.

The variable displacement pump 14 is configured to pump or convey a hydraulic fluid of the hydraulic system 10. For example, the hydraulic fluid is an oil which is also referred to as a hydraulic oil. As will be described in the following, the hydraulic system 10 accomplishes at least two functions: power steering assist and fan motor drive. The variable displacement pump 14 is mechanically driven or drivable by the engine via said through shaft from the air compressor 16 and is capable of being decoupled from the engine by the clutch 20. For example, the hydraulic system 10 comprises an electronic control unit (ECU) configured to control the clutch 20 so that a computer controlled clutching system comprising the ECU and the clutch 20 can be realized. The air compressor 16 is a pneumatic pump, wherein the variable displacement pump 14 is a hydraulic pump. Thus, by means of the computer control clutching system, low and zero efficiency operation of the pneumatic and hydraulic pumps can be minimized. Decoupling the variable displacement pump 14 when not in use reduces parasitic load on the engine, improving the engines efficiency. The variable displacement pump 14 is capable of variable displacement, i.e. pumped volume per revolution. This allows the variable displacement pump 14 to be controlled in either flow control or pressure control regimes, thereby providing several advantages.

The hydraulic system 10 further comprises a first hydraulic circuit 22 through which the hydraulic fluid can flow. At least one hydraulic accumulator 24 is arranged in the first hydraulic circuit 22, the hydraulic accumulator 24 being configured to store the hydraulic fluid. Moreover, the hydraulic system 10 comprises a closed center steering gear 26 of a power steering system, the closed center steering gear 26 being arranged in the first hydraulic circuit 22 so that the steering gear 26 can be supplied with the hydraulic fluid flowing through the first hydraulic circuit 22.

Furthermore, the hydraulic system 10 comprises a second hydraulic circuit 28 through which the hydraulic fluid can flow. At least one hydraulic motor 30 configured as a fan drive motor is arranged in the second hydraulic circuit 28 so that the motor 30 can be supplied with the hydraulic fluid flowing through the second hydraulic circuit 28. Thus, the hydraulic motor 30 is configured to be driven by the hydraulic fluid flowing through the second hydraulic circuit 28. As can be seen from Fig. 1, the variable displacement pump l4is configured to pump the hydraulic fluid to the first and second hydraulic circuits 22 and 28. Furthermore, the hydraulic system 10 comprises a valve device 32 configured to distribute the hydraulic fluid to the first and second hydraulic circuits 22 and 28. Thus, the valve device 32 is configured as a function selection valve as will be described in greater detail in the following.

The hydraulic system 10 comprises a reservoir 34 configured to receive the hydraulic fluid returning from the first and second hydraulic circuits 22 and 28. Furthermore, the hydraulic system 10 comprises a fan 36 configured to convey air, wherein the fan 36 is coupled to the motor 30 so that the fan 36 can be driven by the motor 30 (fan drive motor). Moreover, the hydraulic system 10 comprises a check valve 38 arranged between the hydraulic accumulator 24 and the variable displacement pump l4so as to prevent the hydraulic fluid from flowing from the hydraulic accumulator 24 back to the variable displacement pump 14. Additionally, the hydraulic system 10 comprises at least one cooler 40 configured as an oil cooler. The cooler 40 is arranged in the second hydraulic circuit 28 so that the hydraulic fluid flowing through the second hydraulic circuit 28 can be cooled by means of the cooler 40.

In the following, the flow control and pressure control regimes are described. A flow control regime is used for speed control of the fan 36 (fan speed control), thereby desired fan speed can be maintained independent of the speed of the variable displacement pump 14 (pump speed) and by relation, the engine speed. The hydraulic flow rate is directly proportional to the desired fan drive speed while the system pressure will vary dynamically in response to external loading factors. Variable speed fan control allows the thermal management control strategy to command only the level of cooling power necessary for the current conditions, achieving a reduction in parasitic engine load when compared to a traditional clutched fan drive.

A pressure control regime is used for power steering assist under certain conditions. In this mode, the hydraulic pump displacement is varied to achieve and maintain a desired system pressure. In this way, flow rate varies dynamically based on external factors, but the system pressure is maintained at a constant level. By means of the hydraulic accumulator 24 being a hydraulic pressure accumulator and by means of the variable displacement pump 14 being a variable displacement hydraulic pump an uninterrupted hydraulic power supply to the power steering system and a variable fan speed can be realized in response to cooling demands. Using only one hydraulic pump in the form of the variable displacement pump 14, and only one pneumatic compressor in the form of the air compressor 16 coupled to the same clutch 2015 the least amount of components possible to provide power to the power steering system, power train fan and pneumatic systems. Extremely high voltage and high power density electric motors could eventually replace these systems, but their value may never justify their costs.

The function selection valve (valve device 32) directs the hydraulic fluid to either the hydraulic accumulator 24 and the steering gear 26, or the fan drive motor (motor 30).

Valve direction is electronically controlled by said electronic control unit being a hydraulic system control unit. The function selection valve is what facilitates the merging of multiple hydraulic systems into one, powered by a single pump in the form of the variable displacement pump 14. Additional hydraulic devices could be integrated by adding additional output positions to the function selection valve. The check valve 38 closes off the steering system and prevents back feeding of hydraulic fluid from the hydraulic accumulator 24 to the pump 14, especially while the pump 14 is not active.

The closed center steering gear 26 uses flow only when assist is needed. An orifice opens proportional torsional deflection of the steering valve allowing high pressure fluid to flow at low rates. This allows the high pressure, low volume accumulator to be drawn down relatively slow over time.

A fan drive motor spins the fan 36 and, thus, fan blades of the fan 36 at a speed proportional to the hydraulic flow rate supplied to it. The high-side pressure will change dynamically based on the power required to spin the fan 36, including external factors such as cooling package restrictions, vehicle speed, and fan blade geometry. The internal construction of the fan drive motor includes an overrun mechanism whereby the fan 36 may spin due to windage with a minimal amount of drag, even when there is no hydraulic flow. This is advantageous as the spinning fan 36 presents a lower under-hood flow restriction (better passive cooling) to the cooling package then would be a fixed fan.

The hydraulic accumulator 24 stores a volume of hydraulic fluid under high pressure. This means, the hydraulic accumulator 24 comprises a first chamber 42 configured to store the hydraulic fluid. Moreover, the hydraulic accumulator 24 comprises a second chamber 44 configured to store gas, wherein the chambers 42 and 44 are separated from each other by means of a piston 46 which is translationally movable in a cylinder 48 of the hydraulic accumulator 24. Gas in the separate chamber 44 is pressurized as the hydraulic accumulator 24 or the chamber 42 fills and provides the force necessary to later expel the hydraulic fluid from the chamber 42. The hydraulic accumulator 24 stores energy and delivers it back to the system over time. The control system monitors the pressure level and initiates re-filling of the hydraulic accumulator 24 whenever it is determined to be low.

Fig. 2 shows a diagram illustrating the functionality of the hydraulic accumulator 24. The pre-charged pressure of the accumulator gas determines the minimum hydraulic supply pressure. As the hydraulic accumulator 24 is filled, the gas is compressed into a smaller volume. Because the compression process is nearly isothermal, pressure is inversely proportional to the gas volume. Each time the gas volume is reduced by half, the pressure doubles. Thus, when the accumulator is half full, the pressure (of both the gas and hydraulic fluid) will be twice the pre-charge pressure. At three quarters full, the pressure is four times the pre-charge pressure.

The reservoir 34 provides an expansion volume for the system when the hydraulic accumulator 24 drains. Fluid returning to the reservoir 34 from the fan motor or steering gear 26 is passed through a filter and can be agitated. The reservoir 34 allows the fluid time and space to deaerate before circulating back into the system again. In case of prolonged fan operation in high ambient temperatures, the hydraulic fluid temperature may exceed operational limits. The cooler 40 may optionally be fitted in the hydraulic motor return path to reject system heat to the atmosphere.

In the following, mechanical integration of the hydraulic system 10 will be described. The hydraulic pump 14 is mounted to the air compressor 16 and driven off to the compressor's crankshaft 18. The air compressor 16 is coupled to the engine via the controlled clutch 20. When it is desired to run the pump 14 or the air compressor 16, the clutch 20 is engaged and power is transmitted to both devices. Because of the mechanical coupling both devices will spin when the clutch 20 is engaged, however, it is not always desirable for both devices to be activated or active at the same time, for instance, when the air system is full but the hydraulic accumulator 24 is empty. Therefore, each device must have an unloading method whereby it spins, but does minimal work and consumes minimal energy. For example, in a stand-by mode, both the pump 14 and the air compressor 16 are inactive. In a hydraulic request mode, the pump 14 is active and the air compressor 16 is inactive. In an air request mode the pump 14 is inactive and the air compressor 16 is active and in a hydraulic and air request mode both the pump 14 and the air compressor 16 are active.

The air compressor 16 is unloaded by activating a valve that diverts the air moved by the piston back to the atmosphere at low pressure, instead of into a high pressure storage tank which can be supplied with compressed air by the air compressor 16. The hydraulic pump l4is unloaded by commanding minimum displacement, such that the volume displaced pre-revolution is very small.

In the following, a system control strategy will be explained. The hydraulic system control strategy must generally accomplish two functions. First, monitor the pressure supplied to the steering gear 26 and ensure it is sufficient for the current vehicle operating conditions -activating the hydraulic pump 14 as necessary to fill the hydraulic accumulator 24 or provide constant pressure. Secondly, the system must honor fan speed requests from the thermal management control system by activating the selector valve (valve device 32) and commanding the correct flow rate from the hydraulic pump 14 to achieve the desired fan speed by means of the motor 30. For safety reasons, control of the steering system is prioritized over fan operation.

In the following, a normal refill operation will be described on the basis of Fig. 3. The pressure required to provide adequate steering assistance force varies based on vehicle speed. When stationary and at low vehicle speeds, significant assistance force is required. However, at higher vehicle speeds the required assistance force is much lower.

Therefore, the minimum allowed accumulator pressure is defined as a function of the vehicle speed. Whenever the vehicle is moving, the pressure in the hydraulic accumulator 24 is maintained between the current minimum and maximum pressure levels. When the current pressure is less than or equal to the minimum, the pump 14 is activated and the pressure in the accumulator 24 is increased until the maximum pressure is reached at which point the pump 14 is deactivated.

In the following, a zero-fuel operation will be explained. When the vehicle is slowing or travelling down a sufficiently steep road grade, the engine generates a negative torque on the drive train through internal friction, accessory loads, pumping losses, and in certain conditions, by activating the retarder or engine-brake devices. In this operating mode, no fuel is injected into the engine. The fuel consumption is zero. Under such conditions it can be desirable to add additional (negative) loads to the drive train to increase braking force and take advantage of this braking energy, storing it for later use. Similar to how a hybrid recovers braking energy and stores it as electrical potential in a battery, the hydraulic system 10 can fill and store energy in the hydraulic accumulator 24 during zero-fuel operation.

Zero-fuel hydraulic accumulator filling is enabled when the current engine torque is below a defined torque threshold, the current engine speed is above a defined speed threshold, and the vehicle speed is above a defined speed threshold. Zero-fuel filling is activated when the enable criteria are met and the current accumulator pressure is at or below a defined activation pressure threshold. The zero-fuel filling is deactivated when either the accumulator pressure exceeds a defined deactivation pressure threshold or the enable criteria are no longer met. The defined activation and deactivation pressure thresholds should be assigned values higher and closer together than the normal refill minimum and maximum limits; doing so improves overall energy sufficiency by ensuring an accumulator pressure close to the maximum at the end of the zero-fuel event.

In the following, a top-up detection will be explained. Due to the mechanical integration between the air compressor 16 and the hydraulic pump 14, maximum system efficiency is achieved when both devices are doing useful work. Some energy is wasted when either the air compressor 16 or hydraulic pump 14 must operate in an idle mode while the other is filling. Careful sizing of the air compressor 16, air tanks, hydraulic pump 14, and hydraulic accumulator 24, and of the systems' relative activation periods can improve coordination of the refill events. However, due to the real-world variations in air and hydraulic consumption and differences in each system's filling time, idle mode operation can only be minimized and never eliminated.

Top-up detection logic seeks to minimize activation of the hydraulic pump 14 when the air compressor 16 is not also activated. Filling events for the air system are both long and further spaced than those of the hydraulic system. For example, air tank filling may take seconds every 10 minutes, while accumulator filling only takes a few seconds every 4 minutes. Top-up detections seeks to coordinate the hydraulic accumulator filling with the air tank filling such that both devices reach their full levels at the same moment. Thus, the phasing of the systems' refill periods is synchronized and the number of activations of only the hydraulic system is minimized. This top-up detection is illustrated in Fig. 4. To implement this feature, a model of the air system is created by which the time until the air tanks are full is calculated. In one embodiment, this model may consist of a 2-dimensional look up table that outputs the required tilling time based on inputs of compressor speed and the difference between the current pressure and the cut-out (full tank) pressure.

However, more complex models may also be used to improve accuracy, taking into account other factors such as the current or predicted future air consumption of various vehicle components.

A similar model is implemented for calculating the time required to fill the hydraulic accumulator 24. This model may consist of a similar 2-dimensional look up table with the inputs of hydraulic pump speed and difference between the current and cut-out pressures. When the air system is currently filling the air tank and the remaining time required to fill the air tank is equal to the time required to fill the hydraulic accumulator 24, the hydraulic pump 14 is commanded to fill the accumulator 24. Thus, both systems shall reach their full-fill state at the same moment.

In the following, a stationary operation will be described. When the vehicle is stationary or operated at low speeds, the steering system requires the highest level of power assistance. Rather than activate and deactivate the hydraulic pump 14 to keep the accumulator supplied pressure in a high and narrow range, the pump 14 operates in a pressure-hold mode, maintaining a suitable stationary pressure level which is, for example, 160 bar. If no steering request is detected for sometime, the pump 14 will be deactivated to reduce energy consumption. The pump 14 will return to the pressure-hold mode if steering is again detected. Additionally, this pressure-hold mode shall be disabled when the parking brake is set; a set parking brake indicates that the vehicle is not intended to be operated and a high level of steering assist is unnecessary. In this case, power steering will still be available, but pressure will be allowed to run down to a lower level before being filled back to the stationary pressure level.

Detection of the driver's steering request is not necessarily trivial. Some commercial vehicles are equipped with a steering angle sensor (SAS), typically used as part of the vehicle's advanced stability control system. This system may be configured to provide the current steering wheel angle value to the hydraulic control system (via CAN -Controller Area Network) or other signal interface. In this case, the driver steering demand (turning the wheel) can be inferred directly from observing a sufficient change in current steering wheel angle.

However, in case that this SAS information is not available to the hydraulic system controller (either because it is not fitted or it is not practical to interface the two systems), the driver steering demand can alternatively be inferred through examination of the current hydraulic accumulator pressure and the current hydraulic pump displacement. The algorithm for this inferred detection can be split into two portions: detecting when to activate the pressure-hold mode (start of driver demand) and when to deactivate the pressure-hold mold (end of driver demand).

A condition for activation of pressure-hold mode shall be the detected start of the driver steering demand, which is indicated by a relatively quick drop in accumulator hydraulic pressure. This detection algorithm must be sensitive enough to quickly detect the pressure drop and activate the pressure-hold mode such that the driver maintains consistent and sufficient power assist. Conversely, it must not be so sensitive that the system is falsely triggered due to inherent system pressure drops (internal leakage), measurement noise, or other factors. To this end, the following generalized stability detection filter was conceived to calculate the logical condition signal stable illustrated in Figs. 5 to 7.

(1) Signal_stable shall be true whenever conditional_timer_value »= time_threshold, where time_threshold is a defined parameter.

(2) Conditional timer value shall increment with time whenever signal_delta_condition is true and shall be set to zero whenever a signal delta condition is forced.

(3) Signal_delta_condition shall be true whenever signal delta «= up lini AND signal delta »= down Jim, where up lim and down Jim are defined parameters.

(4) Signal_delta = Input_signal -input_signal_filtered, where input_signal_filtered is a low-pass filtered version of input_signal with a time constant of RC, a defined parameter.

By applying this generalized stability detection filter to the hydraulic accumulator pressure signal, the start of driver steering demand is detected when the calculated stable condition is false, indicating the signal is not stable and dropping rapidly due to the demand from driver steering input. Further improvement in start of demand detection was realized by dynamically adapting the down_lim parameter of the stability detection filter based on the current hydraulic accumulator pressure. This allows the detection sensitivity to be adapted to the accumulator as response characteristics at various pressure levels.

For example, given a certain steering input the pressure drop response at 160 bar is significantly greater than the pressure drop response for the same input at 140 bar.

The activation of the pressure-hold mode shall occur a short time after the end of the driver steering demand is detected, which can be detected by a low and stable hydraulic pump displacement value. The same generalized stability detection filter can also be applied to the pump displacement signal (with separately defined parameters). Thus, the end of a steering demand is detected as follows.

(1) end of demand shall be true whenever stability detection filter (pump displacement) is true AND pump displacement «= DISPL threshold, where DISPL threshold is a defined parameter.

The pressure-hold mode provides a constant high pressure assist to the driver which is necessary for steering maneuvers at low speed. The stationary pressure-hold mode is activated when the vehicle is stationary or moving at low speeds, the park brake is not set, and a driver steering demand is detected as described above. The pressure-hold mode is subsequently deactivated when the vehicle moves faster, the park brake is set, or the end of driver steering demand is detected as described above. Driver steering demand can be detected directly from the SAS unit (if available), or otherwise inferred from the hydraulic accumulator pressure and hydraulic pump displacement signals.

In the following, strategy improvements through prediction are explained. It is well-known from the general prior art to make use of GPS (global positioning system) location and route-based road grade data to predict future vehicle operation. For instance, predictive cruise control can dynamically adjust the target set speed to minimize fuel consumption over an upcoming road section. Similar prediction methods can be applied to improve the energy utilization of the hydraulic system 10.

In the following, predictive zero-fuel alignment will be explained. It is desirable to have the accumulator 24 full at the end of each zero-fuel event because this ensures that the system will go as long as possible before requiring the next activation. This minimizes the parasitic load on the engine during fueling periods, thus, improving fuel efficiency. For example, predictive zero-fuel alignment is illustrated by Fig. 8. Without prediction, an accumulator level close to full at the end of the zero-fuel event can be achieved by setting the zero-fuel pressure activation threshold parameter close to the full pressure level. In this way the accumulator 24 will be guaranteed to end the zero-fuel period with a pressure close to the full pressure level. However, this requires numerous short activations which may have negative consequences on the long-term clutch life and may negatively impact NVH (Noise, Vibration, Harshness).

By combining a longitudinal vehicle dynamics model with a predicted driving profile, the determination of zero-fuel events (no fuel injected into the engine) can be accurately determined in advance. Much as the top-up detection socks to align the hydraulic accumulator filling with the end of the air compressor activation, prediction allows the accumulator filing to be aligned with the end of a zero-fuel event. The prediction model shall calculate the time until the end of the current zero-fuel event. This remaining zero-fuel time is compared to the time currently needed to fill the accumulator 24 as calculated by the hydraulic system model. When the remaining zero-fuel time is equal to the time required to fill the accumulator 24, the hydraulic pump 14 is commanded to fill the accumulator 24. Thus, the accumulator 24 will exit the zero-fuel event completely filled, extending the time before a subsequent activation is required. Thus, the parasitic resistance on the engine during fueling periods is minimized.

In the following, predictive air compressor filling alignment will be explained. The top-up detection discussed above relies on an air system model to calculate the remaining time required to fill the air tanks. Within this model, certain assumptions must be made concerning the future net fill rate which includes the gross filling rate from the compressor 16 (a function of compressor speed and current pressure) as well as the losses from air system consumers (transmission, suspension, etc.). Without prediction, only rough assumptions can be made concerning these factors. For instance, gross filling rate may be calculated given the current engine speed resulting in a certain filling time. However, if some time later a downshift is performed, the engine (and compressor) speed will increase, resulting in an increased gross filling rate and thus a reduced filling time.

The longitudinal vehicle model can be extended to provide more accurate future compressor filling and air consumption rates. Similarly, the hydraulic accumulator filling rate may also depend on the current and future engine speed. This will provide a more accurate filling time calculation and more accurately align the hydraulic accumulator filling with the end of the air compressor filing.

In the following, a fan speed control strategy will be explained. Integration of fan speed control on top of the hydraulic accumulator control creates several unique challenges.

The primary challenge is the fact that the fan circuit requires a flow control regime, while the steering system requires a pressure control regime. One could imagine a specialized pressure and flow split mechanism, however, this complexity is not warranted in practice.

The time required to fill the hydraulic accumulator 24 is so short that the interruption in fan control when using a simple switching valve has negligible effect on thermal management.

In the following, a fan speed control will be explained. The vehicle's thermal management control strategy shall generate a requested fan speed which is provided as an input to the fan speed control strategy which can be seen from Figs. 9 and 10. The goal of the fan speed controller is to modulate the displacement of the hydraulic pump 14 to achieve the requested fan speed via the motor 30. This is achieved by a two-step cascaded RID controller which first generates a displacement request based on current pump speed and difference between the requested and actual fan speed. Using feedback from a displacement sensor, it then generates a physical displacement control signal (voltage or current) which is wired to the physical displacement control mechanism within the hydraulic pump 14.

Fan control with the hydraulic motor 30 and the variable displacement pump 14 provides a benefit over the standard belt-driven on/off and variable speed fan drive systems. Even though a belt drive system is significantly more efficient for overall power transfer, there is a constant parasitic load required to turn the associated belt and pulleys even when the fan is off. With this configuration of a clutched hydraulic pump in the form of the variable displacement pump 14, there is zero energy consumption when the system is inactive.

In the following, fan switching and ramp will be explained on the basis of Fig. 10. In order to provide safe operation, prioritization of the steering system and hydraulic accumulator 24 over the fan drive is required. Filling the hydraulic accumulator 24 during active fan control requires the selector valve to switch from fan control to accumulator filling and back to fan control. During this process it is desirable to avoid sudden changes in hydraulic pressure due to durability and NVH considerations. In the flow control regime of the fan speed control, the system pressure through the selector valve (valve device 32) is relatively low and therefore switching from fan control to accumulator filling can be done with relative ease. However, when filling the hydraulic accumulator 24, system pressure through the selector valve can be very high. Under these conditions, switching the selector valve from accumulator filling to fan control can create a large pressure shock.

Interruption of fan control for an accumulator filling is initiated by directing the selector valve from the motor to the hydraulic accumulator position and the fan speed command is set to zero. To avoid shock while switching back from accumulator filling to fan control the clutch 20 is first opened at the end of filling for a defined time period. This time allows the pump pressure to dissipate from the high levels during filling. After this period, the clutch is activated and the selector valve is directed toward fan operation. Following this, the fan speed command is ramped back to the currently requested level. In conclusion, the following major topics have been discussed and solutions to each problem have been discussed above: -a method for integrating a close-center steering gear 26 and hydraulic accumulator 24 with a hydraulic fan motor 30.

-a method for integrating a variable displacement pump 14 with a clutched air compressor 16 -a method for detecting driver steering demand based on accumulator pressure and pump displacement -a method for managing the pressure of a hydraulic accumulator supplying a close-center steering gear 26 and -a method for optimally coordinating the filling of a hydraulic accumulator 24 and pneumatic storage tank with or without prediction.

List of reference signs hydraulic system 12 drive shaft 14 variable displacement pump 16 air compressor 16 crankshaft clutch 22 first hydraulic circuit 24 hydraulic accumulator 26 closed center steering gear 28 second hydraulic circuit motor 32 valve device 34 reservoir 36 fan 38 check valve cooler 42 chamber 44 chamber 46 piston 48 cylinder

Claims

Claims A hydraulic system (10) for a vehicle, the hydraulic system (10) comprising: -a pump (14) configured to pump a hydraulic fluid and be driven by an engine of the vehicle; -a clutch (20) configured to couple the pump (14) to and decouple the pump (14) from the engine; -a first hydraulic circuit (22) through which the hydraulic fluid can flow; -at least one hydraulic accumulator (24) arranged in the first hydraulic circuit (22), the hydraulic accumulator (24) being configured to store the hydraulic fluid; -a closed center steering gear (26) of a power steering system, the closed center steering gear (26) being arranged in the first hydraulic circuit (22); -a second hydraulic circuit (28) through which the hydraulic fluid can flow; -at least one hydraulic motor (30) arranged in the second hydraulic circuit (28), the hydraulic motor (30) being configured to be driven by the hydraulic fluid flowing through the second hydraulic circuit (28); and -a valve device (32) configured to distribute the hydraulic fluid to the first and second hydraulic circuits (22, 28).
2. The hydraulic system (10) according to claim 1, wherein the pump (14) is configured as a variable displacement pump (14).
3. The hydraulic system (10) according to claim 1 or 2, wherein hydraulic system (10) comprises at least one reservoir (34) configured to receive the hydraulic fluid returning from the first and second hydraulic circuits (22, 28).
4. The hydraulic system (10) according to any one of the preceding claims, wherein hydraulic system (10) comprises a fan (36) configured to convey air, the fan (36) being configured to be driven by the motor (30).
5. The hydraulic system (10) according to any one of the preceding claims, wherein hydraulic system (10) comprises a check valve (38) arranged between the hydraulic accumulator (24) and the pump (14) so as to prevent the hydraulic fluid from flowing from the hydraulic accumulator (24) back to the pump (14).
6. The hydraulic system (10) according to any one of the preceding claims, wherein hydraulic system (10) comprises an air compressor (16) configured to compress air, the air compressor (16) being configured to be driven by the engine.
7. The hydraulic system (10) according to claim 6, wherein pump (14) is configured to be driven by the engine via the clutch (20) and the air compressor (16).
8. The hydraulic system (10) according to any one of the preceding claims, wherein hydraulic system (10) comprises at least one cooler (40) arranged in one of the hydraulic circuits (22, 28), the cooler (40) being configured to cool the hydraulic fluid.
9. A method for operating a hydraulic system (10) of a vehicle, in which method a hydraulic fluid pumped by at least one pump (14) is distributed, by means of a valve device (32), to: -a first hydraulic circuit (22) comprising: o at least one hydraulic accumulator (24) configured to store the hydraulic fluid, and o a closed center steering gear (26) of a power steering system; and -a second hydraulic circuit (28) comprising at least one hydraulic motor (30) being configured to be driven by the hydraulic fluid.
10. A vehicle, in particular a motor vehicle, comprising at least one hydraulic system (10) according to any one of the preceding claims.