Variable turbo supercharger and method of driving the same

A hydraulic servo drive device for driving a slide mechanism of a variable geometry turbocharger includes a servo piston connected to a driveshaft of the slide mechanism and a pilot spool that is accommodated in a center hole of the servo piton and slides by pilot pressure. A first hydraulic chamber and a second hydraulic chamber to and from which pressure oil flows are provided in a housing. The servo piston separately includes a pressure port for introducing pressure oil from an outside, a first piston port for intercommunicating the center hole and the first hydraulic chamber, a second piston port for intercommunicating the center hole and the second hydraulic chamber, and a return port for exiting pressure oil.

This application is a U.S. National Phase Application under 35 USC 371 of International Application PCT/JP2007/068652 filed Sep. 26, 2007.

TECHNICAL FIELD

The present invention relates to a variable geometry turbocharger and a driving method thereof.

BACKGROUND ART

Conventionally, a variable geometry turbocharger in which a nozzle for injecting exhaust gas to an exhaust turbine is structured by a pair of exhaust inlet walls facing each other, one of the pair of exhaust inlet walls advancing toward and retreating away from the other to allow adjustment of a gap between the exhaust inlet walls (i.e., opening area of the nozzle), is known. With the variable geometry turbocharger, at a low speed revolution zone of an engine having a small displacement, the gap between the exhaust inlet walls is reduced to increase a flow speed of exhaust gas flowing into the exhaust turbine, thereby increasing the rotary energy of a turbine wheel to enhance supercharging performance of a charging compressor.

A sliding mechanism having a plurality of links is employed for advancing and retreating the exhaust inlet wall, and the slide mechanism is driven by a pneumatic actuator (e.g., Patent Document 1). Here, the pneumatic actuator is typically structured by a cylinder and a piston that slides within the cylinder, and the piston is slid in a first direction by air pressure of a compressed air and in a second direction by a coil spring biasing the piston theretoward. When the piston is slid in the second direction, supply of the air pressure is interrupted.

Also, employment of a hydraulic servo actuator of the four port type instead of the pneumatic actuator is proposed (e.g., Patent Document 2). According to Patent Document 2, though different from aforementioned Patent Document 1 in the mechanism for achieving a variable opening area of the nozzle, a more precise control of an opening degree can be achieved by driving the variable geometry mechanism by the hydraulic servo actuator. The hydraulic servo actuator switches the supply of the pressure oil to the hydraulic chambers on both sides of the servo piston by a proportional solenoid valve. In other words, a position of a spool forming the solenoid valve is switched to switch the supply of hydraulic pressure to the hydraulic chambers.Patent Document 1: JP-A-11-72008Patent Document 2: JP-T-2003-527522

DISCLOSURE OF THE INVENTION

Problems to be Solved by the Invention

However, according to Patent Document 1, since the piston is reciprocated by different means, i.e., the air pressure and the spring force, a movement of the piston in the first direction is different from a movement of the piston in the second direction, thereby causing difference in movements of the exhaust inlet wall. As a result, hysteresis is increased, making it difficult to precisely control the opening degree of the nozzle. In addition, because a load at the time of sliding the exhaust inlet wall is directly applied on the piston according to the structure, a load drift may be caused depending on largeness of the load, which also hampers precise control of the opening degree. In short, the technique disclosed in Patent Document 1 is an open control technique of the so-called coil balance method, which is not favorable in terms of the hysteresis characteristics and the load drift characteristics.

On the other hand, according to Patent Document 2, the characteristics can be improved by using a hydraulic servo actuator of the four port type. However, according to a structure which switches supply of pressure oil to each hydraulic chamber by a spool of a solenoid valve as disclosed in Patent Document 2: the spool moves in accordance with a balance between a solenoid thrust of the solenoid valve and a spring force of a spring provided within the solenoid valve; a hydraulic circuit opens as a result of the movement of the spool to move a servo piston; a pinion meshing with a rack integrally provided to the servo piston rotates; and an eccentric cam integrated with the pinion rotates to actuate the nozzle opening degree adjustment mechanism. Thus, with this structure, although the spool for controlling the position takes a balance between the solenoid thrust and the spring load, a large amount of pressure oil for driving the servo piston flows through the spool and the spring load is not large enough, so that movement of the spool is likely to be affected by a flow force, thereby limiting preciseness of the spool position control. Incidentally, if the solenoid thrust is increased to increase the spring load, size of the solenoid is increased and a larger space is necessary for the solenoid.

An object of the invention is to provide a variable geometry turbocharger capable of precise control with control characteristics such as the hysteresis characteristic and the load drift characteristic being enhanced and improving reliability, and a driving method of such a variable geometry turbocharger.

Means for Solving the Problems

A variable geometry turbocharger according to an aspect of the invention is a variable turbocharger including: exhaust inlet walls provided at a nozzle at an outer side of a turbine wheel and facing each other; a plurality of nozzle vanes disposed between the exhaust inlet walls with a predetermined interval along a circumferential direction of the turbine wheel; a slide mechanism that advances and retreats one of the exhaust inlet walls in a facing direction relative to the other of the exhaust inlet walls; and a hydraulic servo drive device that drives the slide mechanism, in which the hydraulic servo drive device includes a housing that has an opening at a portion thereof, a servo piston slidably housed in the housing and connected to the slide mechanism via the opening, and a pilot spool that is housed in a center hole of the servo piston and slides by pilot pressure, the housing includes a first hydraulic chamber at a first end of the servo piston and a second hydraulic chamber at a second end of the servo piston, pressure oil being flown in and flown out the first hydraulic chamber and the second hydraulic chamber, the servo piston separately includes a pressure port for introducing the pressure oil from an outside into the center hole, a first piston port for intercommunicating the center hole and the first hydraulic chamber, a second piston port for intercommunicating the center hole and the second hydraulic chamber, and a return port for flowing out the pressure oil of the first and second hydraulic chambers to the outside, and the pilot spool includes a switch that switches an intercommunicating state of the ports.

Incidentally, the switch provided to the pilot spool may be, e.g., a spool land of a pilot spool.

A driving method of a variable geometry turbocharger according to another aspect of the invention is a driving method of the above-mentioned variable geometry turbocharger, the method including: communicating the pressure port with the first piston port and the second piston port with the return port by sliding the pilot spool in a first direction due to increase in pilot pressure, and accordingly making the servo piston follow the sliding of the pilot spool in the first direction; communicating the pressure port with the second piston port and the first piston port with the return port by sliding of the pilot spool in a second direction due to decrease in the pilot pressure, and accordingly making the servo piston follow the sliding of the pilot spool in the second direction; and advancing and retreating the one of the exhaust inlet walls by driving the slide mechanism with sliding of the servo piston.

With the aspects of the invention, because the servo piston and the pilot spool can actualize a hydraulic servo drive device of the four port type, the advancement and retreat of the first exhaust inlet wall can be conducted with a small hysteresis, and the drive load at the time of advancement and retreat is not transmitted to the pilot pool, thus preventing load drift. Accordingly, the control characteristics such as the hysteresis characteristic and the load drift characteristic can be improved, and the opening degree of the nozzle can be controlled with accuracy. In addition, the pilot spool, which functions as the spool of the solenoid valve of Patent Document 2, is operated not by the hydraulic pressure for driving the servo piston but by the pilot pressure independent of this hydraulic pressure. Thus, the pilot spool is prevented from being influenced by flow force, so that the position of the pilot spool can be controlled with more preciseness, thus achieving even more precise control of the opening degree.

Further, because the pilot spool slides within the servo piston, the hydraulic servo drive device can be downsized to prevent enlargement of the variable geometry turbocharger, so that the variable geometry turbocharger can be favorably disposed within a narrow engine room.

In the above arrangement, it is preferable that a pilot hydraulic chamber is provided adjacent to the first end of the servo piston in the housing and partitioned from the first hydraulic chamber by a partition, and the pilot hydraulic chamber is displaced outward in an axial direction of the housing relative to the first hydraulic chamber.

In the above arrangement, it is also preferable that a pilot hydraulic chamber is provided adjacent to the first end of the servo piston in the housing and partitioned from the first hydraulic chamber by a partition, and the pilot hydraulic chamber is displaced inward in a radial direction of the housing relative to the first hydraulic chamber.

With these arrangements, when the pilot hydraulic chamber is provided outwardly in the axial direction of the first hydraulic chamber, radial enlargement of the hydraulic servo drive device can be prevented, and in contrast, when the pilot hydraulic chamber is positioned at a radially inner side of the housing relative to the first hydraulic chamber, the pilot hydraulic chamber and the first hydraulic chamber are radially overlapped to prevent axial enlargement.

In the above arrangement, it is preferable that the servo piston includes a connecting section for connection with the slide mechanism at a position displaced in an axial direction relative to the pressure port.

The pressure port is a portion through which the pressure oil for moving the servo piston passes in a highly pressurized state, so that a shape around the pressure port is likely to influence the movement of the servo piston. Thus, in the aspect of the invention, the connecting section with the slide mechanism is provided at a position apart from the pressure port, so that the shape around the pressure port can be formed to take a favorable hydraulic balance without being affected by the shape of the connecting section, thereby achieving a smooth movement of the servo piston.

In the above arrangement, it is preferable that the slide mechanism includes a converter that converts rotary movement of a rotatable driveshaft into advancing and retreating movement of the one of the exhaust inlet walls, and the driveshaft and the servo piston are connected via a second converter that converts the advancing and retreating movement of the servo piston into rotary movement of the driveshaft.

With this arrangement, a linear movement of the servo piston can be converted into a rotary movement by the converters and again into a linear movement, thus reliably advancing and retreating the exhaust inlet wall.

In the above arrangement, it is preferable that the second converter includes a slide groove formed perpendicular to the axial direction on an outer circumference of the servo piston, a slider that slidably engages in the slide groove, and an arm having a first end rotatably engaged to the slider and a second end connected to the driveshaft.

With this arrangement, the converter, being formed by the slide groove, the slider, and the arm, can be arranged in a simple structure.

In the above arrangement, it is preferable that at least one of the first and second hydraulic chambers is provided with a coil spring that biases the servo piston to one of moving directions of the servo piston.

With this arrangement, because the movement of the servo piston in the first direction is assisted by the coil spring, even when, for some reason, the pressure oil in the piping connected to the hydraulic servo drive device is lost, the spring force of the coil spring can keep the opening degree of the nozzle of the variable geometry turbocharger in a predetermined state.

In the above arrangement, it is preferable that oil for driving the hydraulic servo drive device is lubricating oil of an engine in which the variable geometry turbocharger is installed.

With this arrangement, the oil from the oil pan can be supplied to the hydraulic servo drive device by partly improving the lubrication circuit of the engine, so that the circuit arrangement can be simplified without providing a brand-new hydraulic circuit.

In the above arrangement, it is preferable that the lubricating oil is pressurized and supplied to the hydraulic servo drive device.

With this arrangement, the servo piston of the hydraulic servo drive device can be operated with high hydraulic pressure, so that the slide mechanism can be driven reliably and speedily.

BEST MODE FOR CARRYING OUT THE INVENTION

An embodiment of the invention will be described below with reference to the drawings.

FIG. 1is a cross-sectional view showing a variable geometry turbocharger1according to the embodiment. The variable geometry turbocharger1includes a turbine in a right side ofFIG. 1and a compressor in a left side ofFIG. 1and is provided to an engine body (not shown). A turbine wheel3is housed in a turbine housing2adjacent to the turbine, and a compressor impeller5is housed in a compressor housing4adjacent to the compressor. A shaft6is integrally provided to the turbine wheel3, and the compressor impeller5is attached to an end of the shaft6. The shaft6is rotatably supported by a center housing7. With this arrangement, the turbine wheel3rotated by exhaust gas transmits its rotation to the compressor impeller5via the shaft6, and rotation of the compressor impeller5compresses and charges intake gas.

The turbine housing2is provided with a volute-shaped exhaust inlet path10for introducing exhaust gas from the engine body. The exhaust inlet path10is circumferentially provided continuously with a nozzle11for injecting the exhaust gas toward the turbine wheel3, and the exhaust gas injected from the nozzle11rotates the turbine wheel3before exhausted from an exhaust exit12. The nozzle11is formed by a pair of exhaust inlet walls13and14that face each other.

One exhaust inlet wall13is formed by a lateral side16of an annular movable ring15having a C-shaped cross section. The movable ring15is housed in an annular housing space8provided to the center housing7. A plurality of nozzle vanes17that project toward the other exhaust inlet wall14are attached on the lateral side16of the movable ring15with equal circumferential intervals. A circumferentially continuous recess18is formed on the exhaust inlet wall14, and an end of each nozzle vane17is housed within the recess18. With this structure, when the movable ring15is advanced and retreated by a slide mechanism20described below, the exhaust inlet wall13is moved toward and away from the exhaust inlet wall14to change the opening area of the nozzle11.

Incidentally, because an arrangement of the compressor, which is the same as that of a typical turbocharger, is known, a detailed description thereof will be omitted. The slide mechanism20will be described in detail below.

The slide mechanism20has a structure in which a driveshaft21inserted through a lower side of the center housing7is rotated to advance and retreat the above-noted movable ring15.FIGS. 2 and 3show a primary portion of the slide mechanism20. InFIGS. 2 and 3, a pair of arc-shaped arms22,22extending upward are fixed at intermediate positions of the driveshaft21. A pin23projecting horizontally outward is attached to an end of each arm22, and a slider24is fitted in the pin. The slider24is slidably fitted with a slide groove26adjacent to a base end of a support rod25parallel to the shaft6. A distal end of the support rod25is abutted to a rear side of the movable ring15.

With the above arrangement, when the driveshaft21is rotated, the arm22swings along an axial direction of the shaft6, so that the support rod25is moved to move the movable ring15, thus advancing and retreating the first exhaust inlet wall13relative to the second exhaust inlet wall14. In the slide mechanism20, the support rod25having the arm22, the pin23, the slider24, and the slide groove26forms a converter that converts rotary movement of the driveshaft21to advancing and retreating movement of the exhaust inlet wall13.

The driveshaft21of the slide mechanism20is rotated by a hydraulic servo drive device30via an arm27provided on an end of the driveshaft21. The hydraulic servo drive device30will be described in detail below.

As shown inFIG. 4, a basic structure of the hydraulic servo drive device30is rotating the driveshaft21as result of vertical reciprocation of a servo piston31. Thus, a slide groove32perpendicular to an axial direction is provided on an outer circumference of the servo piston31; a pin28projecting toward the slide groove32is provided on the arm27adjacent to the driveshaft21; a slider29is fitted in the pin28; and the slider29is slidably fitted with the slide groove32.

In other words, in the embodiment, another converter, which includes the slide groove32, the slider29, the pin28, and the arm27, is provided for converting the reciprocating movement of the servo piston31into the rotary movement of the driveshaft21. With the vertical movement of the servo piston31, the slider29moves up and down and slides along the slide groove32, and the movement of the slider29and the rotation of the pin28allow an arc movement of the arm27to rotate the arm27.

FIG. 5shows a vertical cross section of the hydraulic servo drive device30. InFIG. 5, the hydraulic servo drive device30includes: the servo piston31; a housing33which slidably houses this servo piston31and a portion of which forms an opening33A; and a pilot spool36which is housed in a center hole34axially penetrating the servo piston31and slides by pilot pressure. The hydraulic servo drive device30is mounted in the center housing7of the variable geometry turbocharger1via an O-ring100that seals a surrounding of the opening33A.

The housing33, which has a prismatic external shape, contains a vertically penetrating cylindrical cylinder space35in inside thereof, and the servo piston31is housed in the cylinder space35. Upper and lower ends of the cylinder space35are hermetically covered by covers37and38via the O-rings101and102. A connecting section39of the driveshaft21and the servo piston31is formed at a position adjacent to the opening33A of the housing33. Thus, the size of the opening33A is determined in consideration of sliding amount of the servo piston31and the slider29.

A side of the housing33remote from the opening33A includes: a pilot port41for supplying pilot pressure from, e.g., a proportional solenoid valve95(FIG. 8) positioned apart from the variable geometry turbocharger1; a pump port42for supplying pressure oil from a pressure elevation pump92(FIG. 8); and a drain port43for returning the pressure oil. The pressure elevation pump92and the proportional solenoid valve95are installed in the same engine body (not shown) as the one in which the variable geometry turbocharger1of the embodiment is installed. Because the proportional solenoid valve95is provided to the engine body independently of the housing33, the housing33can be downsized, so that the variable geometry turbocharger1itself can be downsized to save space. Such a space saving advantage is important for a construction machine or the like that has an extraordinarily small engine room unlike a transport truck or the like.

The cylinder space35of the housing33is partitioned by a partition44into a portion where the servo piston31slides and a portion thereabove. The partition44abuts to a stepped portion formed on an inner circumference of the cylinder space35, and an O-ring103for sealing the space partitioned by the partition44is provided in the vicinity of the abutting portion. The partition44is provided with a tubular portion45extending downward, and the tubular portion45is inserted in an upper side of the center hole34of the servo piston31. The upper one of the spaces partitioned by the partition44forms a pilot hydraulic chamber46, which is communicated with the pilot port41.

On the other hand, the lower one of the spaces partitioned by the partition44forms a first hydraulic chamber47which is defined by the partition44and an upper end of the servo piston31. In other words, the pilot hydraulic chamber46is displaced axially outward (upward in the embodiment), thereby preventing enlargement of the hydraulic servo drive device30as a whole. In addition, a second hydraulic chamber48is formed between a lower end of the servo piston31and the lower cover38.

Next, the servo piston31will be described. The servo piston31is provided with a pressure port51for intercommunicating the center hole34and the pump port42of the housing33and for delivering the pressure oil from the pump into the center hole34. Outer sides of the pressure port51are opened grooves in formed radially opposing to each other, and since the grooves have a predetermined vertical dimension, the pressure port51and the pump port42are constantly communicated in the strokes of the servo piston31.

In addition, the servo piston31is provided with a return port52that intercommunicates the center hole34and the drain port43of the housing33to return the pressure oil in the center hole34to a tank. An outer side of the return port52is opened in a groove formed on an outer circumference of the servo piston31, so that the return port52and the drain port43are also constantly communicated in the strokes of the servo piston31. Also, in the embodiment, since the connecting section39of the servo piston31and the driveshaft21is provided at a position opposite to the return port52, the connecting section39is displaced downward in the axial direction relative to the pressure port51.

As shown inFIG. 5by dotted lines, the servo piston31is further provided with a first piston port53for intercommunicating the center hole34and the upper first hydraulic chamber47and a second piston port54for intercommunicating the center hole34and the lower second hydraulic chamber48. Here, the opening of the first piston port53adjacent to the center hole34is positioned more downward than the opening of the pressure port51, and the opening of the second piston port54adjacent to the center hole34is positioned more upward than the opening of the pressure port51. The first and second piston ports53and54are each displaced so as not to communicate with the pressure port51or the return port52.

An abutment member55is screwed with the servo piston31via an O-ring104to hermetically close the lower side of the center hole34. The servo piston31abuts to the cover38via the abutment member55, and abutment position serves as the lowermost position of the servo piston31. A coil spring56is disposed between the cover38and the abutment member55within the second hydraulic chamber48to assist an upward movement of the servo piston31. Even if the pressure oil in piping to the hydraulic servo drive device30is lost due to, e.g., a trouble of the pressure elevation pump92, spring force of the coil spring56keeps the nozzle opening degree of the variable geometry turbocharger1at a rather opened state (preferably at a fully opened state).

The pilot spool36includes two spool lands, i.e., first and second spool lands61and62(switch of the invention) at a substantially central portion thereof. A return flow path63opened downward is provided to an inside of the pilot spool36. An upper groove of the first spool land61and the return flow path63are communicated while a lower groove of the second spool land62and the return flow path63are also communicated. In addition, since the lower side of the return flow path63is opened, this return flow path63, the return port52, and the drain port43are communicated.

The pilot spool36is vertically slidable in the center hole34of the servo piston31through the tubular portion45of the partition44, and an upper end of the pilot spool36is screwed and fixed to a holder64disposed within the pilot hydraulic chamber46. The holder64is biased upward by a coil spring65in the pilot hydraulic chamber46. The pilot spool36is moved downward by pilot pressure resisting the biasing force of the coil spring65and upward by the biasing force of the coil spring65with return of the pilot pressure oil (drained to an oil pan80adjacent to the solenoid valve95though the drain flow path is not shown).

In the hydraulic servo drive device30having such an arrangement, when the pilot spool36is elevated relative to the servo piston31, the servo piston31follows the elevation, and when the pilot spool36is lowered, the servo piston31follows the lowering movement. Here, since the pilot spool36only slides axially in the servo piston31, drive load at the time of advancement and retreat of the movable ring15is applied on the servo piston31via the slide mechanism20but not at all on the pilot spool36.

Accordingly, when position of the pilot spool36is controlled for position control of the servo piston31and further for advancing and retreating the movable ring15to change the opening area of the nozzle11, the position control of the pilot spool36can be conducted without being influenced by the drive load, so that load drift can be eliminated. Thus, even when fluid pressure deriving from exhaust gas is unstable in a turbocharger, that is, even in a case of the variable geometry turbocharger1of the embodiment, the opening area of the nozzle11can be easily controlled for precise control of emission. In addition, because position control can be precisely conducted, control format may be changed from the feedback control to the feedforward control to reduce response time and to handle transients with accuracy.

Next, operation of the hydraulic servo drive device30will be specifically described with reference toFIGS. 5 to 7. InFIG. 5, because the pilot pressure that overcomes the biasing force of the coil spring65is supplied, both the pilot spool36and the servo piston31are at a lowermost position. Thus, in this state, a lower end of the pilot spool36abuts to an upper end of the abutment member55, and a lower end of the abutment member55abuts to the cover38. Further, at this position, the upper spool land61of the pilot spool36is displaced downward relative to the second piston port54; the second piston port54is communicated with the return port52through the return flow path63; and the pressure oil in the second hydraulic chamber48is drained.

On the other hand, the lower second spool land62is also displaced downward relative to the first piston port53, and the pressure port51and the first piston port53are communicated. Accordingly, the pressure oil is supplied to the first hydraulic chamber47through the pressure port51and the first piston port53.

Incidentally, a portion of the pressure oil supplied to the pilot hydraulic chamber46passes through a slight gap formed between the tubular portion45of the partition44and the holder64or a slight gap formed between the tubular portion45and an outer circumference of an upper end of the pilot spool36, and enters a space defined therebelow, that is, a space defined by an inner circumference of the center hole34of the servo piston31, an outer circumference of the pilot spool36, and a lower end of the tubular portion45.

When the pilot pressure is lowered from this state to a predetermined value by returning the pressure oil of the pilot hydraulic chamber46as shown inFIG. 6, the pilot spool36is elevated to a position where the pilot pressure is balanced with the force of the coil spring65. At this time, the upper first spool land61is displaced to an upper side of the second piston port54, so that the second piston port54and the pressure port51become communicated to supply the pressure oil to the second hydraulic chamber48.

At the same time, because the lower second spool land62is also displaced to an upper side of the first piston port53, the first piston port53and the return flow path63become communicated, and a portion of the pressure oil in the first hydraulic chamber47is drained, so that the servo piston31follows the elevation of the pilot spool36. This elevation of the servo piston31ends when the first and second piston ports53and54are closed by the first and second spool lands61and62, and the servo piston31pauses at a position corresponding to the position where the pilot spool36pauses. The servo piston31does not go past the pilot spool36during the elevation.

Next, as shown inFIG. 7, when the pilot pressure is completely released, the pilot spool36moves upward to a position where an upper end of the holder64abuts to a ceiling of the pilot hydraulic chamber46, and the servo piston31following this movement elevates until the upper end thereof abuts to the partition44. At this time, the pilot spool36and the servo piston31are both at an uppermost position, and the first and second piston ports53and54are respectively closed by the first and second spool lands61and62with the second hydraulic chamber48full of the pressure oil.

Here, the pressure oil that has entered the space defined by the inner circumference of the center hole34of the servo piston31, the outer circumference of the pilot spool36, and the lower end of the tubular portion45returns to the pilot hydraulic chamber46through the above-mentioned gap.

When the servo piston31is to be lowered to a predetermined position, the pilot pressure is supplied to lower the pilot spool36to a predetermined position. With this operation, the second piston port54is again communicated with the return flow path63to drain a portion of the pressure oil of the second hydraulic chamber48, thus lowering the servo piston31. This lowering movement ends when the first and second piston ports53and54are closed by the first and second spool lands61and62, and the servo piston31pauses at a position corresponding to the position where the pilot spool36pauses. The servo piston31does not go past the pilot spool36during the lowering movement.

With the hydraulic servo drive device30which operates as described above, the servo piston31and the pilot spool36function as a four-port valve of the triple position type, so that both the upward movement and the downward movement of the servo piston31can be conducted by supply of the pressure oil to one of the first and second hydraulic chambers47and48and drain of the pressure oil from the other occurring simultaneously with the supply. Thus, the hysteresis characteristic can be greatly improved as compared with the conventional open control of the spring balance type. Accordingly, because the load drift does not occur and the hysteresis characteristic is favorable, adjustment of the opening degree of the nozzle11can be precisely conducted. Further, because the pilot spool36operates not by solenoid thrust but by pilot pressure, unlike Patent Document 2, the pilot spool36is not affected by the flow force of the pressure oil, thereby achieving more precise position control of the pilot spool36.

In addition, the pilot spool36for switching the supply of the pressure oil to the first and second hydraulic chambers47and48also has a function that corresponds to the spool of the solenoid valve of Patent Document 2. The arrangement where this pilot spool36slides within the servo piston31contributes to downsizing of the hydraulic servo drive device30, thereby preventing enlargement of the variable geometry turbocharger1. Moreover, although the embodiment requires such a solenoid valve as in Patent Document 2 for supplying pilot pressure, such a solenoid valve can be disposed at any suitable position apart from the variable geometry turbocharger1to lessen heat influence, so that a malfunction at the solenoid valve can be prevented, thus enhancing reliability.

FIG. 8schematically shows a lubrication circuit70of an engine in which the variable geometry turbocharger1of the embodiment is installed. In the lubrication circuit70, the lubricating oil in the oil pan80is pumped up by a hydraulic pump81and supplied to a main gallery84via an oil cooler82and an oil filter83. The lubricating oil from the main gallery84mainly lubricates a crankshaft85and a camshaft86.

The lubrication circuit70includes the following paths that are branched from the main gallery84: an injector-side path71for lubricating a cam driver or the like in a fuel injector87; a transmission-mechanism-side path72for lubricating a power transmission mechanism88that includes a timing gear; a rocker-arm-side path73for lubricating a rocker arm89; a turbocharger-side path74for lubricating a bearing portion that supports the shaft6of the variable geometry turbocharger1; and a first drain path75for returning the lubricating oil from the variable geometry turbocharger1and the fuel injector87to the oil pan80. In addition, in the embodiment, a pressure oil supply path90for supplying a portion of the lubricating oil to the hydraulic servo drive device30as the driving pressure oil and a second drain path91for returning the pressure oil to the oil pan80from the drain port43of the hydraulic servo drive device30are provided separately from the lubrication circuit70.

In other words, in the embodiment where the pressure oil for driving the hydraulic servo drive device30is fed by a portion of an engine lubricating oil, the path for supplying the pressure oil is the pressure oil supply path90branched before the main gallery84. The pressure elevation pump92is provided adjacent to a base end of the pressure oil supply path90, and the pressurized pressure oil is supplied to the pump port42of the hydraulic servo drive device30through a driving pressure path93adjacent to a distal end of the pressure oil supply path90. A discharge pressure of the hydraulic pump81is approximately in the range of 196 to 294 kN/m2(2 to 3 kg/cm2), and a discharge pressure after pressurization by the pressure elevation pump92is approximately 1470 kN/m2(15 kg/cm2). Here, the distal end of the pressure oil supply path90is branched into the driving pressure path93for supplying the pump port42and a pilot pressure path94for supplying pilot pressure to the pilot port41of the hydraulic servo drive device30, and thus, the pilot pressure path94is provided with the proportional solenoid valve95for generating the pilot pressure. By applying a predetermined electric current to the solenoid valve95, pilot pressure in the range of 0 to 1470 kN/m2(0 to 15 kg/cm2) corresponding to the electric current can be generated to move the pilot spool36to a position corresponding to the pilot pressure.

Incidentally, although the best arrangement, method, and the like for carrying out the invention have been described above, the scope of the invention is not limited thereto. In other words, although a particular embodiment of the invention is mainly illustrated and described, a variety of modifications may be made by those skilled in the art on shapes, amounts, and other detailed arrangements of the embodiment set forth above without departing from the scope of the inventive idea and the object of the invention.

Accordingly, the above description limiting shapes, amounts and the like is exemplary description for facilitating understanding of the invention and does not limit the scope of the invention, so that description with names of members without all of or a portion of the limitations such as limitations on shapes or amounts are included in the scope of the invention.

With this structure, since the hydraulic chambers46and47are aligned with each other, and an axial dimension of the housing33can be reduced, thereby further facilitating downsizing of the hydraulic servo drive device30.

The invention can be utilized as a variable geometry turbocharger, e.g., for a construction machine that has a narrow engine room and is typically equipped with a hydraulic pump.