US20130202452A1

US20130202452A1 - Hydraulic fan drive

Info

Publication number: US20130202452A1
Application number: US13/638,016
Authority: US
Inventors: Michael Schuette; Tobias Pfruender; Martin Fassbender; Egon Rill
Original assignee: Robert Bosch GmbH
Current assignee: Robert Bosch GmbH
Priority date: 2010-03-30
Filing date: 2011-03-03
Publication date: 2013-08-08
Also published as: EP2553231A2; CN102812218B; WO2011120486A2; JP2013524107A; CN102812218A; DE102011012905A1; EP2553231B1; KR20130018253A; WO2011120486A3; JP5599504B2

Abstract

A hydraulic fan drive includes a hydraulic pump that has an adjustable swept volume. The hydraulic pump is assigned a pressure control valve arrangement that is configured to regulate a pump pressure by adjustment of the swept volume. The drive further includes a hydraulic motor configured to drive an impeller wheel and a pressure line, which is connected to a pressure input of the hydraulic motor and into which pressure medium is configured to be conveyed by the hydraulic pump. A hydraulic accumulator is connected to the pressure line and the displacement of the hydraulic motor is configured to be adjusted. Energy is buffer-stored by the hydraulic accumulator by feeding in pressure medium beyond the amount that is displaced by the hydraulic motor. This stored energy becomes free for other operations of a machine, for example, to which the drive is attached.

Description

The invention starts from a hydraulic fan drive, which has a hydraulic pump of adjustable swept volume, which is assigned a pressure control valve arrangement for regulating a pump pressure by adjustment of the swept volume, a hydraulic motor for driving a fan impeller, and a pressure line, which is connected to a pressure inlet of the hydraulic motor and into which pressure medium can be delivered by the hydraulic pump. Fan drives of this kind can be used especially in construction machines, agricultural and forestry machines, in conveying applications, in trucks and buses and in rail vehicles.
A fan drive of this kind is known from DE 43 21 637 A1, for example. In this document, the pressure-regulated hydraulic pump is operated with a hydraulic motor of constant displacement in an open hydraulic circuit. The pressure control valve arrangement consists essentially of a control valve having a pressure port, which is connected to the pressure outlet of the hydraulic pump and at which therefore the pump pressure is available, a tank port connected to a tank, and a control port connected to the adjusting chamber at an adjusting piston, of a directly controlled pressure limiting valve that can be adjusted by electroproportional means, and of a nozzle, which is arranged between the pressure outlet of the hydraulic pump and the inlet of the pressure limiting valve. The control piston of the control valve is acted upon by the pump pressure so as to produce a fluidic connection of the pressure outlet of the hydraulic pump and so as to reduce the swept volume (delivery volume per revolution) of the hydraulic pump with the control port and is acted upon by a spring and by the pressure prevailing at the inlet of the pressure limiting valve so as to produce a fluidic connection of the control port to the tank port and so as to increase the displacement. A particular setting of the pressure limiting valve thus gives a particular pump pressure and therefore a particular torque at the hydraulic motor and hence a particular speed of the fan impeller.
It is the underlying object of the invention to develop a hydraulic fan drive having the features from the preamble of patent claim 1 in such a way that it can be used to recover energy.
According to the invention, this object is achieved by virtue of the fact that, in the hydraulic fan drive having the features from the preamble of patent claim 1, a hydraulic accumulator is connected to the pressure line, and the hydraulic motor has an adjustable displacement. In the case of a hydraulic fan drive according to the invention, connecting the hydraulic accumulator to the pressure line makes it possible to buffer-store energy by feeding in pressure medium beyond the amount which is displaced by the hydraulic motor, this energy being released in other operations on the machine, for example during a braking operation or during the lowering of a load. The pressure changes in the pressure line which are associated with the buffer storage and the output of energy can be compensated for in such a way by a change in the displacement of the hydraulic motor that the torque which is output by the hydraulic motor corresponds to the desired fan speed. The hydraulic accumulator is preferably connected directly to the pressure line without valves, which require actuation.
Advantageous embodiments of a hydraulic fan drive according to the invention can be found in the dependent claims.
Additional energy can be fed in directly by the hydraulic pump, for example, if, in accordance with patent claim 2, the setting of the pressure control valve arrangement can be varied by remote control. In normal operation, it is advantageous if the pressure control valve arrangement is set to a pressure which lies midway between the maximum and the minimum accumulator operating pressure. If the pressure control valve arrangement is then set to a higher pressure, additional pressure medium can be fed into the hydraulic accumulator in the event of a braking operation or a sudden relief of the load on a diesel engine driving the hydraulic pump, for example, in order to provide a speed safety feature for said engine. The increased pressure level is compensated for by a reduction in the displacement of the hydraulic motor, ensuring that there is no change in the fan speed.
If, as indicated in patent claim 3, a check valve, which closes toward the hydraulic pump, is arranged between the hydraulic pump and a section of the pressure line to which the hydraulic accumulator is connected, pressure medium can be fed into the pressure line independently of the hydraulic pump and independently of the pressure setting thereof to a higher pressure level.
According to patent claim 4, it is then also possible to feed pressure medium in from an external pressure medium source via a branch line that opens into the pressure line. According to patent claim 5, it is advantageous if there is a check valve opening toward the section of the pressure line in the branch line. Thus, the pressure in the branch line upstream of the check valve can also fall below the pressure in the pressure line or can even be tank pressure.
The buffer-stored energy can be used to drive the hydraulic motor and, where there is a check valve at the pump outlet, an adjustable pressure control valve arrangement can be set to the original value again immediately after the feeding in of an additional volume of pressure medium. Without a check valve, it may be advantageous if the pressure control valve arrangement is in each case set to a somewhat higher pressure than the instantaneous accumulator pressure. This can also be achieved by means of a slow time-dependent withdrawal of the control signal for the pressure control valve arrangement.
This is particularly expedient if, in accordance with patent claim 6, the hydraulic pump can also be operated as a hydraulic motor while retaining the direction of rotation. In principle, this is already possible with the aid of a directional control valve, by means of which the high-pressure port and the low-pressure port of the hydraulic pump can be interchanged.
However, it appears more advantageous if, in accordance with patent claim 7, the pump is a hydraulic pump that can be adjusted via zero, and it can therefore also be operated as a hydraulic motor with the same pressure port and the same direction of rotation. In this case, there is advantageously no check valve arranged between the pressure port of the hydraulic pump and the hydraulic accumulator. By setting the pressure control valve arrangement to a lower pressure than that prevailing in the hydraulic accumulator, the hydraulic pump is made to swivel via zero and to operate as a hydraulic motor. The resulting fall in the pressure level in the hydraulic accumulator is compensated for in respect of the fan speed by increasing the displacement of the hydraulic motor. The fan impeller continues to rotate at the desired speed. In this motor mode, the hydraulic pump acting as a motor supports the diesel engine.
By charging the hydraulic accumulator when less driving power is required from the diesel engine and discharging the hydraulic accumulator via the hydraulic pump, which then operates as a hydraulic motor, when there is a high demand for driving power, the power output by the diesel engine can be smoothed out or held constant. The pressure control valve arrangement can also be set to values below the normal pressure level, thus making available a particularly large amount of energy to support the diesel engine or, more generally, an internal combustion engine or even an electric motor (primary unit). As an option, the hydraulic motor can even be set to a displacement of zero so that, although the fan impeller is not driven for a brief period, all the energy stored is available to support the primary unit.
Even if pressure medium is fed into the pressure line and hence into the hydraulic accumulator by some pressure medium source other than the hydraulic pump, the primary unit is subject to less stress by virtue of more uniform loading and, as a result, has a more advantageous consumption of primary energy.
According to the particularly preferred embodiment in patent claim 8, the hydraulic motor is assigned a torque control valve arrangement for regulating a motor torque by adjusting the displacement. Given a particular control signal, the hydraulic motor is then in each case set to the displacement which gives the torque corresponding to the desired speed of the fan impeller at the pressure currently prevailing in the pressure line and in the hydraulic accumulator. This is automatically compensated in the case of pressure fluctuations. By changing a hydraulic control pressure or an electric control signal, the torque characteristic can be shifted in parallel. A particularly simple parallel arrangement of a plurality of torque-regulated fan motors to drive a plurality of fan impellers is also possible.
The torque control valve arrangement is preferably designed in accordance with patent claim 9.
According to patent claim 10, the hydraulic motor can also be assigned a control valve arrangement, by means of which the displacement of the hydraulic motor can be varied proportionally to a control signal, wherein the control signal is dependent, on the one hand, on a setpoint speed value of the hydraulic motor and, on the other hand, on a detected speed of the hydraulic motor or on the detected pressure in the pressure line. From the detected speed, it is possible to ascertain directly, by comparison with the setpoint speed, whether the displacement has to be increased or reduced. From the detected pressure and the setpoint speed, the setpoint displacement can be calculated and specified.

Three illustrative embodiments of a hydraulic fan drive according to the invention are shown in the drawings. The invention will now be explained in greater detail with reference to the figures of these drawings, in which:

FIG. 1 shows the first illustrative embodiment, in which the hydraulic pump has a remote controllable pressure control valve arrangement for varying the pressure level in a hydraulic accumulator and has a swept volume that can be adjusted via zero,

FIG. 2 shows the second illustrative embodiment, in which the hydraulic pump has a pressure control valve arrangement set to a fixed value and the pressure level in the hydraulic accumulator can be raised by means of an external pressure medium source and the hydraulic motor is torque-controlled, and

FIG. 3 shows the third illustrative embodiment, in which, as in the first illustrative embodiment, the hydraulic pump has a remote controllable pressure control valve arrangement for varying the pressure level in a hydraulic accumulator and has a swept volume that can be adjusted via zero and in which, as in the second illustrative embodiment, the hydraulic motor is torque-controlled.

According to FIG. 1, the hydraulic fan drive comprises a first hydraulic machine 10, which is so called because it can be operated both as a hydraulic pump and as a hydraulic motor. The hydraulic machine 10 is connected mechanically to a diesel engine 11. It has a high-pressure port (pressure port) 12 and a low-pressure port (tank port) 13, which is continuously connected to a tank 9. From the pressure port 12, a pressure line 14 leads to a hydraulic motor 15, by which a fan impeller 16 can be driven. Connected directly to the pressure line is a hydraulic accumulator 17, which can be operated in a pressure range of between 100 bar and 300 bar, for example, and is charged to 200 bar in normal operation.
The hydraulic machine is, for example, an axial piston machine of swashplate construction, the swept volume of which can be adjusted via zero between a maximum positive value and a maximum negative value. In the case of a positive swept volume, the hydraulic machine 10 delivers pressure medium into the pressure line 14 as a hydraulic pump. In the case of a negative swept volume, the hydraulic machine operates as a hydraulic motor in the same direction of rotation and is supplied with pressure medium from the pressure line 14.
For adjustment of the swept volume, there are two adjusting pistons and a spring, of which one adjusting piston 18, which has a larger effective area than the other adjusting piston, delimits an adjusting chamber 19, to which pressure medium can be fed and from which pressure medium can be released.
This feeding and discharge of pressure medium is controlled by a pressure control valve arrangement 20 mounted on the hydraulic machine 10. This arrangement includes a continuously adjustable control valve 21 with a zero overlap or a small positive overlap between the control edges and with a pressure port 22, which is connected fluidically to the pressure port 12 of the hydraulic machine 10, a tank port 23, which is connected to a leakage oil port 30 via the interior of the housing of the hydraulic machine 10, and a control port 24, which is connected via a damping nozzle 25 to the adjusting chamber 19. The valve piston of the control valve is acted upon by the pump pressure so as to connect the control port 24 to the pressure port 22 and so as to reduce the swept volume of the hydraulic machine 12 as far as negative values and by a spring 26 and a variable control pressure so as to connect the control port 24 to the tank port 23 and so as to increase the swept volume, said control pressure being picked off between a control oil nozzle 27 and a pressure limiting valve 28 that can be adjusted proportionally by an electromagnet 29, i.e. corresponding to the set value of the pressure limiting valve. If the control valve is in the control position in which the valve piston makes at most small movements about a central position, the control oil flow is always the same, irrespective of the pressure level, since the pressure in the pressure port of the hydraulic machine is then always higher by the pressure equivalent of the spring 26 than the control pressure, and hence the pressure difference across the nozzle 27 is always the same. The pressure limiting valve 28 has a falling characteristic, i.e. the pressure at the inlet thereof is all the lower, the greater the level of energization of the electromagnet 29. As a consequence, the pressure limiting valve has its maximum set value if the electric system fails, and the pressure at the pressure port of the hydraulic machine 12 accordingly rises to a maximum.
The hydraulic motor 15 is preferably one of axial piston construction, in particular of oblique axis construction, and can be adjusted between a displacement of zero and a maximum displacement. It is connected to the pressure line 14 by a pressure port 35 and to tank 9 by a tank port 36. For variation of the displacement, the hydraulic motor 15 has an adjusting piston 37, which is provided on one side with a piston rod 38 and thus separates an annular space 39 on the piston rod side and an adjusting chamber 40 on the opposite side from the piston rod. The inflow and outflow of pressure medium to and from the adjusting chamber 40 is controlled by a control valve 42, which can be acted upon in a proportional manner by an electromagnet 41, is mounted on the hydraulic motor 15 and has a pressure port 43, which is connected to pressure port 35, a tank port 44, which is connected to tank 9 via the interior of the housing of the hydraulic motor and a leakage port (not designated specifically), and a control port 45, which is connected to the adjusting chamber 40. The annular space 39 has a continuous fluidic connection to pressure port 35 via the housing (not designated specifically) of the control valve 42.
The electromagnet 41 acts on a control piston of the control valve 42 so as to connect the control port 45 to pressure port 43 and so as to reduce the displacement of the hydraulic motor 15. So as to connect the control port to the tank port 44, the control piston is acted upon by a first spring 46, which is supported in a fixed location relative to the housing and by means of which the start of control can be set, and by a second spring 47, which is arranged between the control piston and the piston rod 38, i.e. the adjusting piston 37. This construction has the effect that the position of the adjusting piston and hence the displacement of the hydraulic motor 15 depends directly on the force of the electromagnet 41, i.e. on the level of current flowing through the coil of the electromagnet.
In a compensated state, the adjusting piston must namely, on the one hand, be at rest and, on the other hand, there must be an equilibrium of forces on the control piston of the control valve, which is in the control position thereof, irrespective of the position of the adjusting piston. The sum of the forces exerted by the two springs 46 and 47 must therefore be equal to the force of the electromagnet 41. In the control position of the control piston, the force of the spring is always the same. Therefore, the force of the spring 47 must differ according to the magnetic force. This differing force of the spring 47 results from the differing positions of the adjusting piston 37, depending on the magnetic force. This kind of adjustment is also known as electroproportional adjustment.
The speed of the hydraulic motor 15 and hence that of the fan impeller 16 is detected by a speed sensor 50, which sends a corresponding signal to an electric control unit 51. The latter is furthermore supplied with a setpoint speed value, which is determined from the temperature of a medium to be cooled. The control unit 51 then controls the electromagnet 41 in such a way that the desired speed is achieved on the basis of the displacement that is established.
As an alternative to a speed sensor, it is also possible to provide a pressure sensor 52, by means of which the pressure in the pressure line 14 is detected. In the case of a fan impeller, there is a fixed relationship between the driving torque and the speed. The pressure can thus be used to calculate the displacement that is necessary to produce the driving torque required to reach or maintain the desired speed, and the electromagnet can be controlled accordingly.
In pure fan mode, the hydraulic machine 10 operates as a hydraulic pump and is set to a pressure value of 150 bar, for example. This pressure prevails in the pressure line 14 and in the hydraulic accumulator 17. The displacement of the hydraulic motor 15 is then set in such a way that, at the pressure of 150 bar, the driving torque required to drive the fan impeller at the desired speed is obtained. If the electric system fails, the hydraulic machine is set to maximum pressure and the hydraulic motor is set to maximum displacement, thus ensuring adequate cooling of the medium to be cooled in all cases.
The setting of the pressure value at the hydraulic machine 10 and hence the state of charge of the hydraulic accumulator 17 can be varied in order, for example, to enable the hydraulic motor 15 to be operated predominantly with a large displacement or to smooth the power to be output by the diesel engine 11. If the diesel engine is underused at any particular time, the pressure level can be raised briefly while, if it is overloaded, the pressure level can be lowered briefly. By raising the pressure level, it is possible to protect the diesel engine from overspeeding.
In the case of a braking operation, the braking energy can be used to enable the hydraulic machine 10 to be driven as a hydraulic pump by the vehicle itself. In this case, the pressure can be set to a maximum value, with the result that the hydraulic machine 10 adjusts to a maximum swivel angle and the braking effect is the product of the maximum swept volume and the instantaneous pressure in the hydraulic accumulator.
In the event of drooping of the diesel, appropriate energization of the electromagnet 29 is used to set a pressure which is lower than the pressure in the hydraulic accumulator, with the result that the hydraulic machine swivels via zero to the negative swept volume of maximum magnitude and, operating as a hydraulic motor, supports the diesel engine until the accumulator pressure has fallen to the pressure set at the hydraulic machine 10.
Braking energy previously stored in the hydraulic accumulator can thus be used to boost the diesel engine. However, it can also be used to supply the hydraulic motor 15 and hence to drive the fan impeller. In that case, the pressure set at the hydraulic machine should be lowered sufficiently slowly, taking into account the quantity of pressure medium consumed by the hydraulic motor 15, that the set pressure is not lower than the accumulator pressure.
The hydraulic fan drive shown in FIG. 2 includes a hydraulic pump 60, which can be connected mechanically to a primary unit 61 by a clutch and can be driven by said unit. No motor mode is provided for the machine 60 in this case. Accordingly, the swept volume of the hydraulic pump can be adjusted only between a minimum value close to or equal to zero and a maximum value. The hydraulic pump has a pressure port 62 and a suction port 63, which is continuously connected to a tank 9. A pressure line 14 leads from the pressure port 62 to a hydraulic motor 65, by which a fan impeller 16 can be driven. Once again, a hydraulic accumulator 17 is connected to the pressure line and can be operated in a pressure range between 100 bar and 300 bar, for example. There is a check valve 66 in the pressure line 14 between the hydraulic accumulator and the pressure port 62, said check valve blocking flow toward the hydraulic pump 60.
The hydraulic pump 60 is an axial piston pump of swashplate construction, for example.
To adjust the swept volume of the hydraulic pump 60, there are, as in the hydraulic machine 10 from the first illustrative embodiment, two adjusting pistons, of which one adjusting piston 18, which has a larger effective area than the other adjusting piston, delimits an adjusting chamber 19, to which pressure medium can be fed and from which pressure medium can be released.
This feeding and discharge of pressure medium is controlled by a pressure control valve arrangement 70 mounted on the hydraulic pump 60. This arrangement includes a continuously adjustable control valve 71 with a zero overlap or a small positive overlap between the control edges and with a pressure port 72, which is connected fluidically to the pressure port 62 of the hydraulic pump 60, a tank port 73, which is connected to a leakage oil port 67 via the interior of the housing of the hydraulic pump 60, and a control port 74, which is connected to the adjusting chamber 19. The valve piston of the control valve is acted upon by the pump pressure so as to connect the control port 74 to the pressure port 62 and so as to reduce the swept volume of the hydraulic pump 60 and exclusively by a spring 75 so as to connect the control port 74 to the tank port 73 and so as to increase the swept volume. During operation, the pressure established at the pressure port 62 of the hydraulic pump 60 is in each case the pressure equivalent of the spring 75, e.g. a pressure of 100 bar. In contrast to the illustrative embodiment in FIG. 1, remote adjustment of this pressure is not provided. However, the setting of the spring 75 can be changed in the course of commissioning or during servicing work.
The hydraulic motor 65 is preferably one of axial piston construction, in particular of oblique axis construction, and, like the hydraulic motor 15 in the illustrative embodiment shown in FIG. 1, can be adjusted between a displacement of zero and a maximum displacement. It is connected to the pressure line 14 by a pressure port 76 and to tank 9 by a tank port 77. For variation of the displacement, the hydraulic motor 15 has an adjusting piston 78, which is provided on one side with a piston rod 79 and thus separates an annular space 80 on the piston rod side and an adjusting chamber 81 on the opposite side from the piston rod. The inflow and outflow of pressure medium to and from the adjusting chamber 81 is controlled by a torque control valve arrangement 69 having a control valve 82, which is mounted on the hydraulic motor 65 and has a pressure port 83, which is connected to pressure port 76, a tank port 84, which is connected to tank 9 via the interior of the housing of the hydraulic motor and a leakage port (not designated specifically), and a control port 85, which is connected to the adjusting chamber 81. The annular space 80 has a continuous fluidic connection to pressure port 76.
A control piston of the control valve 82 is acted upon by a spring 86 supported in a fixed location relative to the housing so as to connect the control port 85 to the tank port 84 and so as to increase the displacement of the hydraulic motor 65, and is acted upon via a control line 87 by a variable control pressure, by means of which a force that can be varied by remote control can thus be exerted on the control piston. The control pressure is advantageously at a maximum if the electric system fails. So as to connect the control port to the pressure port 83, the control piston is acted upon by a force which depends on the position of the adjusting piston 78 and hence on the displacement of the hydraulic motor 65 and on the pressure in the pressure line 14. For this purpose, the control piston is initially supported on a lever 88 at a distance which is constant in the control position from an axis of rotation of said lever 88, said axis being fixed relative to the housing. Conversely, the lever 88 is acted upon by way of a rod 89, which is inserted in a movable manner in the adjusting piston 78 and on which the pressure prevailing in the pressure line 14 acts. The torque produced at the lever 88 by way of the rod thus represents the product of the pressure prevailing at the hydraulic motor 65 and the displacement of the hydraulic motor and hence the output torque of the hydraulic motor. An opposing torque on the lever 88 is produced by the sum of the forces exerted by the spring 86 and the control pressure on the control piston of the control valve 82. In a stable state, the sum of the torques acting on the lever 88 must be zero. If the pressure in the pressure line 14 rises, for example, the torque exerted on the lever 88 by way of the rod 89 becomes greater than the torque exerted by way of the control piston. The lever is turned and the control piston is moved, thus connecting the control port 85 to the pressure port 83 of the control valve 82. Pressure medium flows to the adjusting chamber 81, and the adjusting piston moves in a direction corresponding to a reduction in the displacement. The rod 89 moves with the adjusting piston along the lever 88, with the result that the lever arm for the pressure force acting via the rod becomes smaller until there is once again an equilibrium between the torques. Thus, while there is no change in the control pressure in the control line 87, the torque exerted by the hydraulic motor 65 does not change when there is a change in the pressure level in the pressure line 14. Conversely, the torque and hence the fan speed can be varied by changing the control pressure.
In the illustrative embodiment shown in FIG. 2, provision is then made to enable pressure medium to be fed into the hydraulic fan circuit independently of the hydraulic pump 60. In this case, the hydraulic accumulator 17 makes it possible for the quantity fed in to be larger than the instantaneous quantity being consumed by the hydraulic motor 65, without the excess quantity being ejected via a pressure limiting valve.
Two possibilities for an additional infeed are shown in FIG. 2. In addition to hydraulic pump 60, there is another hydraulic pump 90, said pump being adjustable in an electroproportional manner, for example, and being able to be coupled to the drive train of the vehicle by means of a clutch 91. In the event of a braking operation, the clutch is closed, and the hydraulic pump 90 is driven by the vehicle and delivers pressure medium into the pressure line 14 via a check valve 92, which closes in the direction of said pump. The advantage here is that it is always possible to feed in pressure medium, irrespective of the pressure in the fan circuit.
The second possibility shown is that of feeding in pressure medium when a load is being lowered, said load being indicated here by a hydraulic cylinder 95. To control the lowering movement, a flow control valve 96 with a proportionally adjustable metering orifice (not shown specifically) and a pressure compensator arranged in series therewith is provided. In a branch line 94 between the flow control valve and the pressure line 14 there is a check valve 97, which blocks flow toward the flow control valve. A 2/2-way valve 98, by means of which a flow to the tank can be opened, is connected to the fluidic connection between the flow control valve 96 and the check valve 97.
At the desired lowering speed, pressure medium displaced from the hydraulic cylinder can be fed into the pressure line 14 if the load pressure is higher than the pressure in the pressure line 14 and in the hydraulic accumulator 17 by the pressure drop across the metering orifice of the flow control valve 96. It may be that this is the situation at the beginning of a lowering movement but that it does not continue since the pressure in the hydraulic accumulator 17 rises. The valve 98 must then be opened. Whether it is still possible to feed in pressure medium can be determined with the aid of pressure sensors which detect the load pressure and the accumulator pressure. Position monitoring of the pressure compensator is also possible. When the pressure compensator is fully open, it is no longer possible to feed in pressure medium. However, there is a loss of information in this case after the opening of the directional control valve 98, making it impossible to ascertain whether the accumulator pressure has fallen back below the load pressure after a certain time because of the pressure medium consumption of the hydraulic motor 65. It is only during the next lowering operation that feeding in pressure medium can be attempted again.
With respect to the additional quantity of pressure medium fed in, it is possible to distinguish between two cases. If the quantity of pressure medium fed in is less than the consumption of the hydraulic motor 65, the residual quantity is delivered by the hydraulic pump 60, and the pressure level in the system remains at the level set at the hydraulic pump 60. In this case too, an energy saving is achieved since the hydraulic pump 60 swivels to a smaller swept volume.
If the additional quantity of pressure medium fed in is greater than the consumption of the hydraulic motor 65, the pressure in the fan circuit rises beyond the level set at the hydraulic pump 60. The hydraulic pump is adjusted back to zero stroke by cutting off the pressure. The pressure level can then assume a significantly higher value, e.g. 300 bar, until it is limited by a pressure limiting valve 99 connected to the pressure line 14. If no more or only a smaller quantity is fed in than the hydraulic motor 65 is consuming, the hydraulic accumulator 17 initially takes over the supply of the fan circuit completely or partially until the hydraulic pump 60 swivels back out at a pressure level of 100 bar.
In principle, electronic control of the output torque of the hydraulic motor 65 is also conceivable in the illustrative embodiment shown in FIG. 2, the accumulator pressure being detected by a sensor and the swept volume being set to match the necessary torque.
In the illustrative embodiment shown in FIG. 2, an anti-cavitation valve 100 that opens from the tank port to the pressure port 76 is depicted between the pressure port 76 and the tank port 77. There can also be such an anti-cavitation valve in the illustrative embodiment shown in FIG. 1.
In the fan drive shown in FIG. 3, a hydraulic pump in accordance with the illustrative embodiment shown in FIG. 1 and a hydraulic motor in accordance with the illustrative embodiment shown in FIG. 2 are combined. The reference numbers from FIGS. 1 and 2 are therefore used for corresponding parts in FIG. 3. The fan drive shown in FIG. 3 thus includes a first hydraulic machine 10, which is so called because it can be operated both as a hydraulic pump and as a hydraulic motor. The hydraulic machine 10 is connected mechanically to a diesel engine 11. It has a high-pressure port (pressure port) 12 and a low-pressure port (tank port) 13, which is continuously connected to a tank 9. From the pressure port 12, a pressure line 14 leads to a hydraulic motor 65, by which a fan impeller can be driven. Connected directly to the pressure line is a hydraulic accumulator 17, which can be operated in a pressure range of between 100 bar and 300 bar, for example, and is charged to 200 bar in normal operation. A pressure limiting valve 99 is also connected to the pressure line 14.
The hydraulic machine 10 is, for example, an axial piston machine of swashplate construction, the swept volume of which can be adjusted via zero between a maximum positive value and a maximum negative value. In the case of a positive swept volume, the hydraulic machine 10 delivers pressure medium into the pressure line 14 as a hydraulic pump. In the case of a negative swept volume, the hydraulic machine operates as a hydraulic motor in the same direction of rotation and is supplied with pressure medium from the pressure line 14.
For adjustment of the swept volume, there are two adjusting pistons and a spring, of which one adjusting piston 18, which has a larger effective area than the other adjusting piston, delimits an adjusting chamber 19, to which pressure medium can be fed and from which pressure medium can be released.
This feeding and discharge of pressure medium is controlled by a pressure control valve arrangement 20 mounted on the hydraulic machine 10. This arrangement includes a continuously adjustable control valve 21 with a zero overlap or a small positive overlap between the control edges and with a pressure port 22, which is connected fluidically to the pressure port 12 of the hydraulic machine 10, a tank port 23, which is connected to a leakage oil port 30 via the interior of the housing of the hydraulic machine 10, and a control port 24, which is connected via a damping nozzle 25 to the adjusting chamber 19. The valve piston of the control valve is acted upon by the pump pressure so as to connect the control port 24 to the pressure port 22 and so as to reduce the swept volume of the hydraulic machine 12 as far as negative values and by a spring 26 and a variable control pressure so as to connect the control port 24 to the tank port 23 and so as to increase the swept volume, said control pressure being picked off between a control oil nozzle 27 and a pressure limiting valve 28 that can be adjusted proportionally by an electromagnet 29, i.e. corresponding to the set value of the pressure limiting valve. If the control valve is in the control position in which the valve piston makes at most small movements about a central position, the control oil flow is always the same, irrespective of the pressure level, since the pressure in the pressure port of the hydraulic machine is then always higher by the pressure equivalent of the spring 26 than the control pressure, and hence the pressure difference across the nozzle 27 is always the same. The pressure limiting valve 28 has a falling characteristic, i.e. the pressure at the inlet thereof is all the lower, the greater the level of energization of the electromagnet 29. As a consequence, the pressure limiting valve has its maximum set value if the electric system fails, and the pressure at the pressure port of the hydraulic machine 12 accordingly rises to a maximum.
The hydraulic motor 65 in the illustrative embodiment shown in FIG. 3 is preferably one of axial piston construction, in particular of oblique axis construction, and, like the hydraulic motor 15 in the illustrative embodiment shown in FIG. 1, can be adjusted between a displacement of zero and a maximum displacement. It is connected to the pressure line 14 by a pressure port 76 and to tank 9 by a tank port 77. For variation of the displacement, the hydraulic motor 15 has an adjusting piston 78, which is provided on one side with a piston rod 79 and thus separates an annular space 80 on the piston rod side and an adjusting chamber 81 on the opposite side from the piston rod. The inflow and outflow of pressure medium to and from the adjusting chamber 81 is controlled by a torque control valve arrangement 69 having a control valve 82, which is mounted on the hydraulic motor 65 and has a pressure port 83, which is connected to pressure port 76, a tank port 84, which is connected to tank 9 via the interior of the housing of the hydraulic motor and a leakage port (not designated specifically), and a control port 85, which is connected to the adjusting chamber 81. The annular space 80 has a continuous fluidic connection to pressure port 76.
A control piston of the control valve 82 is acted upon by a spring 86 supported in a fixed location relative to the housing so as to connect the control port 85 to the tank port 84 and so as to increase the displacement of the hydraulic motor 65, and is acted upon via a control line 87 by a variable control pressure, by means of which a force that can be varied by remote control can thus be exerted on the control piston. The control pressure is advantageously at a maximum if the electric system fails. So as to connect the control port to the pressure port 83, the control piston is acted upon by a force which depends on the position of the adjusting piston 78 and hence on the displacement of the hydraulic motor 65 and on the pressure in the pressure line 14. For this purpose, the control piston is initially supported on a lever 88 at a distance which is constant in the control position from an axis of rotation of said lever 88, said axis being fixed relative to the housing. Conversely, the lever 88 is acted upon by way of a rod 89, which is inserted in a movable manner in the adjusting piston 78 and on which the pressure prevailing in the pressure line 14 acts. The torque produced at the lever 88 by way of the rod thus represents the product of the pressure prevailing at the hydraulic motor 65 and the displacement of the hydraulic motor and hence the output torque of the hydraulic motor. An opposing torque on the lever 88 is produced by the sum of the forces exerted by the spring 86 and the control pressure on the control piston of the control valve 82. In a stable state, the sum of the torques acting on the lever 88 must be zero. If the pressure in the pressure line 14 rises, for example, the torque exerted on the lever 88 by way of the rod 89 becomes greater than the torque exerted by way of the control piston. The lever is turned and the control piston is moved, thus connecting the control port 85 to the pressure port 83 of the control valve 82. Pressure medium flows to the adjusting chamber 81, and the adjusting piston moves in a direction corresponding to a reduction in the displacement. The rod 89 moves with the adjusting piston along the lever 88, with the result that the lever arm for the pressure force acting via the rod becomes smaller until there is once again an equilibrium between the torques. Thus, while there is no change in the control pressure in the control line 87, the torque exerted by the hydraulic motor 65 does not change when there is a change in the pressure level in the pressure line 14. Conversely, the torque and hence the fan speed can be varied by changing the control pressure.
In the illustrative embodiment shown in FIG. 3, as in the illustrative embodiment shown in FIG. 2, provision is made to enable pressure medium to be fed into the hydraulic fan circuit independently of the hydraulic pump 10. In this case, the hydraulic accumulator 17 makes it possible for the quantity fed in to be larger than the instantaneous quantity being consumed by the hydraulic motor 65, without the excess quantity being ejected via a pressure limiting valve.
In the illustrative embodiment shown in FIG. 3, the feeding of pressure medium into the pressure line and into the hydraulic accumulator 17 when a load is being lowered is shown, said load being indicated here by a hydraulic cylinder 95. To control the lowering movement, a flow control valve 96 with a proportionally adjustable metering orifice (not shown specifically) and a pressure compensator arranged in series therewith is provided. In a branch line 94 between the flow control valve and the pressure line 14 there is a check valve 97, which blocks flow toward the flow control valve. A 2/2-way valve 98, by means of which a flow to the tank can be opened, is connected to the fluidic connection between the flow control valve 96 and the check valve 97.
At the desired lowering speed, pressure medium displaced from the hydraulic cylinder can be fed into the pressure line 14 if the load pressure is higher than the pressure in the pressure line 14 and in the hydraulic accumulator 17 by the pressure drop across the metering orifice of the flow control valve 96. It may be that this is the situation at the beginning of a lowering movement but that it does not continue since the pressure in the hydraulic accumulator 17 rises. The valve 98 must then be opened. Whether it is still possible to feed in pressure medium can be determined with the aid of pressure sensors which detect the load pressure and the accumulator pressure. Position monitoring of the pressure compensator is also possible. When the pressure compensator is fully open, it is no longer possible to feed in pressure medium. However, there is a loss of information in this case after the opening of the directional control valve 98, making it impossible to ascertain whether the accumulator pressure has fallen back below the load pressure after a certain time because of the pressure medium consumption of the hydraulic motor 65. It is only during the next lowering operation that feeding in pressure medium can be attempted again.
In the illustrative embodiment shown in FIG. 3 too, as in the illustrative embodiment shown in FIG. 2, provision can be made to feed in pressure medium by means of an additional hydraulic pump, which can be driven by the diesel engine 11.
With respect to the additional quantity of pressure medium fed in, it is possible to distinguish between two cases. If the quantity of pressure medium fed in is less than the consumption of the hydraulic motor 65, the residual quantity is delivered by the hydraulic pump 60, and the pressure level in the system remains at the level set at the hydraulic pump 60. In this case too, an energy saving is achieved since the hydraulic pump 60 swivels to a smaller swept volume.
If the additional quantity of pressure medium fed in is greater than the consumption of the hydraulic motor 65, the pressure in the fan circuit rises beyond the level set at the hydraulic pump 60. The hydraulic pump is adjusted back to zero stroke by cutting off the pressure, it being possible here in addition to employ a swivel angle sensor, by means of which the swivel angle is sensed. The valve 28 would then be set in each case in such a way that the swivel angle is zero or just above zero. The pressure level can then assume a significantly higher value, e.g. 300 bar, until it is limited by a pressure limiting valve 99 connected to the pressure line 14. If no more or only a smaller quantity is fed in than the hydraulic motor 65 is consuming, the hydraulic accumulator 17 initially takes over the supply of the fan circuit completely or partially until the hydraulic pump 60 swivels back out at a pressure level of 100 bar.
In the illustrative embodiment shown in FIG. 3 too, an anti-cavitation valve 100 that opens from the tank port to the pressure port 76 is arranged between the pressure port 76 and the tank port 77 of the hydraulic motor 65.
In principle, electronic control of the output torque of the hydraulic motor 65 is also conceivable in the illustrative embodiment shown in FIG. 3, the accumulator pressure being detected by a sensor and the swept volume being set to match the necessary torque.
In the illustrative embodiment shown in FIG. 3, it is thus possible to use an external energy source, e.g. a lifting cylinder in a lowering operation or an additional hydraulic pump to drive the fan impeller. Moreover, brief support of the diesel engine is possible by means of the hydraulic machine 10.
In a fan drive according to the invention, in particular in a fan drive with EP adjustment of the hydraulic motor in accordance with the illustrative embodiment shown in FIG. 1 or electric adjustment with swivel angle restoration with the aid of an electric swivel angle sensor, it is also possible to use a hydraulic motor that can be swiveled via zero. In this case, the direction of rotation of the fan impeller can be reversed without an additional valve in order to blow the radiator clean. This is advantageous particularly in the case of forestry and construction machines.

Claims

1. A hydraulic fan drive, comprising:

a hydraulic pump of defining an adjustable swept volume, the hydraulic pump being assigned a pressure control valve arrangement configured to regulate a pump pressure by adjustment of the swept volume,

a hydraulic motor configured to drive a fan impeller, the hydraulic motor having an adjustable displacement,

a pressure line, which is connected to a pressure port of the hydraulic motor and into which pressure medium is configured to be delivered by the hydraulic pump and

a hydraulic accumulator connected to the pressure line.

2. The hydraulic fan drive as claimed in claim 1, wherein the setting of the pressure control valve arrangement is configured to be varied by remote control.

3. The hydraulic fan drive as claimed in claim 1, further comprising a check valve arranged between the hydraulic pump and a section of the pressure line to which the hydraulic accumulator is connected, the check valve being configured to close toward the hydraulic pump.

4. The hydraulic fan drive as claimed in claim 1, further comprising a branch line, via which pressure medium is configured to be fed into the pressure line independently of the hydraulic pump, the branch line opening into the pressure line.

5. The hydraulic fan drive as claimed in claim 4, further comprising a check valve opening toward the section of the pressure line in the branch line.

6. The hydraulic fan drive as claimed in claim 1, wherein the hydraulic pump is configured to be operated as a hydraulic motor while retaining the direction of rotation.

7. The hydraulic fan drive as claimed in claim 6, wherein the hydraulic pump configured to be adjusted via zero and is further configured to be operated as a hydraulic motor with the same pressure port and the same direction of rotation.

8. The hydraulic fan drive as claimed in claim 1, wherein the hydraulic motor is assigned a torque control valve arrangement configured to regulate a motor torque by adjusting the displacement.

9. The hydraulic fan drive as claimed in claim 8, wherein the torque control valve arrangement has a valve piston which is acted upon by a force that is configured to be varied by remote control so as to increase the displacement of the hydraulic motor and is supported at an at least approximately fixed distance from an axis of rotation on a lever, wherein a pressure force dependent on the pressure in the pressure line and acting on the valve piston so as to reduce the displacement acts on the lever in the opposite direction, and wherein the distance of the point of introduction of the pressure force into the lever from the axis of rotation thereof depends on the position of an adjusting piston of the hydraulic motor and on the displacement of the hydraulic motor.

10. The hydraulic fan drive as claimed in claim 1, wherein the hydraulic motor is assigned a control valve arrangement by which the displacement of the hydraulic motor is configured to be varied proportionally to a control signal, and wherein the control signal is dependent, on a setpoint speed value of the hydraulic motor and on a detected speed of the hydraulic motor or on the detected pressure in the pressure line and in the hydraulic accumulator.

11. The hydraulic fan drive as claimed in claim 1, wherein the hydraulic accumulator is connected directly to the pressure line.