NL2012449B1

NL2012449B1 - Seven-ton class hydraulic excavator.

Info

Publication number: NL2012449B1
Application number: NL2012449A
Authority: NL
Inventors: Tsuruga Yasutaka; Tanaka Nozomu; Manabe Takeo; Mizoguchi Kazuhiko; Kuno Makoto; Nakayama Yoshimi; Takahashi Kiwamu; Mori Kazushige; Yoshida Hajime; Matsui Takuya
Original assignee: Hitachi Construction Mach Co
Priority date: 2013-03-29
Filing date: 2014-03-17
Publication date: 2016-08-01
Also published as: JP2014198937A; JP5809657B2; NL2012449A; DE102014104347A1

Abstract

Output power of an engine 1 is set to be 36 kW, lower than current 40 to 55 kW output power. Consumption (absorption) horsepower of a hydraulic pump 2 is set to be substantially the same as current consumption horsepower of a hydraulic pump usually mounted on a seven-ton class hydraulic excavator based on related art. The conduits 4a, 4b, Sa, 5b extending from a control valve 3 to a boom cylinder 4 and an arm cylinder 5 are configured to have a size of 3/4 inch (1.91 cm) which is larger than a current hose size adopted in the related art. The conduits 7a, 7b extending to traveling motors 7, 8 are configured to have a size of 5/8 inch (1.59 cm) which is larger than a current hose size adopted in the related art. As a result, a hydraulic excavator that, despite its seven-ton class specifications, can be equipped with the engine free of structurally complex and expensive engine peripheral devices, and that maintains performance comparable to that of a conventional working machine is provided.

Description

SEVEN-TON CLASS HYDRAULIC EXCAVATOR BACKGROUND OF THE INVENTION 1. Field of the Invention

The present invention relates to a seven-ton class hydraulic excavator. 2. Description of the Related Art

Some of hydraulic excavators are called seven-ton class hydraulic excavators. In general, the seven-ton class hydraulic excavators refer to hydraulic excavators of the specifications where they are equipped with a boom of a single type instead of an offset type, and a bucket as an attachment and have a vehicle weight of 6,300 to 8,500 kg, a bucket size of 0.28 to 0.33 m3, a maximum excavation radius of 6,300 to 7,700 mm, a maximum excavation depth of 3,800 to 4,700 mm, and a maximum excavation height of 6,700 to 7,800 mm.

Gas emission regulations relating to diesel engines have been reinforced year after year, and these gas emission regulations vary from country to country.

In the EU, the gas emission regulations on hydraulic excavators and other non-road motor vehicles for special purposes are provided for in Directive 2004/26/EC (Non-Patent Document 1) and Directive 2010/26/EU (Non-Patent Document 2). According to these provisions, the gas emission regulations relating to engines of an output power range of 37<P<56, where P is engine output power, have been tightened from Stage IIIA to Stage NIB since 2013. For engines of an output power range of 18<P<37, however, gas emission regulations Stage IIIA remain effective after 2013.

RELATED ART

Non-Patent Document 1: Directive 2004/26/EC Non-Patent Document 2: Directive 2010/26/EU

Summary of the Invention

Conventional seven-ton class hydraulic excavators are equipped with a 40-55 kW engine. This engine output power falls into the 37<P<56 output power range under the gas emission regulations. These engines have been required to comply with the Stage NIB of the gas emission regulations in EU since 2013, and therefore, there are needed new peripheral devices for reducing the content of exhaust gas constituents (NOx and PM) to an even lower level than before. More specifically, the excavators should be installed with, for example, a common rail system (CRS), an exhaust gas recirculation (EGR) valve, and a diesel particulate filter (DPF), which make the engine structure complex and hence will increase costs. Loading them boosts a risk of a failure occurring in any of the engine peripherals such as CRS, EGR, and DPF.

An object of the present invention is to provide a seven-ton class hydraulic excavator that, despite its seven-ton class specifications, can be equipped with an engine free of structurally complex and expensive engine peripheral devices, and that maintains performance comparable to that of a conventional working machine.

In order to solve the foregoing problems, the seven-ton class hydraulic excavator according to the present invention is configured in such a manner that output power of the engine is set to be 34.0 to 36.4 kW, lower than current 40 to 55 kW output power of seven-ton class hydraulic excavators. The conduits connecting a control valve to a boom cylinder and an arm cylinder are configured to have a size larger than 1/2 to 5/8 inch (1.27 to 1.59 cm): current conduit sizes of the seven-ton class hydraulic excavators.

The engine output power of 34.0 to 36.4 kW is classified into the 18<P<37 output power range under the EU gas emission regulations, and the EU gas emission regulations relevant to this value fall under Stage IIIA, one level lower than Stage NIB. From 2013 onward, therefore, the engine further continues to meet the EU gas emission regulations with the use of existing peripheral devices.

In addition, when the engine output power is set to be 34.0 to 36.4 kW, lower than the current 40 to 55 kW output power, the engine output power decreases in margin ratio: a ratio of excess pump consumption horsepower over the engine output horsepower to the engine output power. However, there are not many highlands and other areas where it reaches high temperatures in the EU, and limiting the usage environment to the inside the EU enables setting of a margin ratio lower than a general value. Furthermore, the hydraulic pump has a horsepower control function added to keep the consumption (absorption) horsepower below a set level. The setting of the margin ratio is based on the absorption of an excessive hydraulic pump load caused by a delay in dynamic response of the horsepower control function when the hydraulic pump steeply changes in delivery pressure. The present invention focuses attention upon the fact that the steep change in the delivery pressure of the hydraulic pump is caused mainly by changes in excavation load pressures of the boom cylinder and the arm cylinder. The conduits connecting the control valve to the boom cylinder and the arm cylinder are thus configured to have a size larger than 1/2 to 5/8 inch (1.27 to 1.59 cm), which is the current conduit size of the seven-ton class hydraulic excavators. This reduces an energy loss due to a conduit pressure loss, corrects the decrease in margin ratio, and thus enables the excavator to maintain the performance comparable to that of a conventional working machine even in the eventuality of the decline in margin ratio.

The conduits connecting the control valve to the boom cylinder and the arm cylinder are preferably configured to have a size of 3/4 inch (1.91 cm).

In addition, the output power of the engine is preferably 35 kW.

Furthermore, the plurality of hydraulic actuators in the present invention includes left and right traveling motors that drive left and right crawler belts, respectively. The conduits connecting the control valve to the left and right traveling motors are preferably configured to have a size larger than 1/2 inch (1.27 cm), current conduit sizes of the seven-ton class hydraulic excavators. The decrease in margin ratio is corrected as is the case with the boom cylinder and the arm cylinder even when the load pressure on traveling motors instantaneously increases. As a result, the excavator can develop the performance comparable to that of a conventional working machine even in the eventuality of the decrease in margin ratio.

The conduits connecting the control valve to the left and right traveling motors are preferably configured to have a size of 5/8 inch (1.59 cm).

Moreover, the consumption horsepower of the hydraulic pump is preferably set to be substantially equal to a current consumption horsepower level of a hydraulic pump usually mounted on the seven-ton class hydraulic excavators.

According to the present invention, a seven-ton class hydraulic excavator can employ an engine clearing the EU gas emission regulations at the Stage MIA level, dispensing with complex peripheral devices (CRS, EGR, DPF), and suppressing a rise in costs. This makes it possible to maintain performance comparable to that of a conventional working machine even in the eventuality of a reduction in margin ratio.

BRIEF DESCRIPTION OF THE DRAWINGS

Figure 1 is a system block diagram showing a hydraulic drive system of a seven-ton class hydraulic excavator according to an embodiment of the present invention;

Figure 2 is a torque control diagram of a torque control device (first and second torque control pistons);

Figure 3 is an external view of the seven-ton class hydraulic excavator;

Figure 4 is a diagram that lists general specifications of a conventional seven-ton class hydraulic excavator in a tabular format;

Figure 5 is a diagram that lists gas emission regulations of various countries;

Figure 6 is a diagram that lists boom cylinder, arm cylinder, and traveling motor conduit sizes of the present embodiment in a tabular format;

Figure 7 is a graph showing a change in an excess of pump consumption horsepower over engine output horsepower between working machines in related art and the present embodiment (the present invention);

Figure 8 is a diagram that shows changes with time in the engine output horsepower and pump consumption horsepower of the working machine in the related art; and

Figure 9 is a diagram that shows changes with time in the engine output horsepower and pump consumption horsepower of the working machine in the present embodiment (the present invention).

DESCRIPTION OF THE PREFERRED EMBODIMENTS

An embodiment of the present invention will hereunder be described. - Configuration -

Figure 1 is a system block diagram showing a hydraulic drive system of a seven-ton class hydraulic excavator according to an embodiment of the present invention.

The hydraulic drive system of the seven-ton class hydraulic excavator is equipped with a diesel engine 1, a hydraulic pump 2 driven by the diesel engine 1, and a plurality of hydraulic actuators including a boom cylinder 4, an arm cylinder 5, a bucket cylinder 6, a right traveling motor 7, a left traveling motor 8, a swinging motor 9, a blade cylinder 10, and a swing cylinder 11, each driven by a hydraulic fluid delivered from the hydraulic pump 2. The hydraulic drive system further has a control valve 3 that incorporates a plurality of spools for controlling a flow of the hydraulic fluid supplied from the hydraulic pump 2 to the hydraulic actuators, and a return circuit 13 that is connected to the control valve 3 and that causes a return fluid from the hydraulic actuators to circulate back into a drain tank 12.

The hydraulic pump 2 is a split-flow pump including two delivery ports P1 and P2, and one capacity control mechanism (swash plate) 2a. The delivery ports P1, P2 are connected to the control valve 3 via supply lines 14 and 15, and the control valve 3 is connected to the actuators 4 to 11 via respective lines 4a, 4b, 5a, 5b, 6a, 6b, 7a, 7b, 8a, 8b, 9a, 9b, 10a, 10b, 11a, 11b.

The control valve 3 is constructed such that when the boom cylinder 4 is being driven, streams of the fluid that has been delivered from the delivery ports P1, P2 of the hydraulic pump 2 merge into one stream which is then supplied to the boom cylinder 4. Similarly, when the arm cylinder 5 is being driven, the streams of the fluid delivered from the delivery ports P1, P2 of the hydraulic pump 2 merge into one stream is then supplied to the arm cylinder 5. The control valve 3 is also constructed in such a manner that when the right traveling motor 7 is being driven, the fluid that has been delivered only from the delivery port P1, one of the two delivery ports, of the hydraulic pump 2 is supplied to the right traveling motor 7. When the left traveling motor 8 is being driven, the fluid that has been delivered only from the other delivery port P2 of the hydraulic pump 2 is supplied to the right traveling motor 7. The control valve 3 is further constructed such that when any of the hydraulic actuators 6 and 9 to 11, other than the boom cylinder 4, the arm cylinder 5, the right traveling motor 7, and the left traveling motor 8, is being driven, the fluid that has been delivered from either of the two delivery ports P1 and P2 of the hydraulic pump 2 is supplied to the actuator.

The hydraulic pump 2 includes a pump control unit 21 that limits a flow rate of the hydraulic fluid delivered from the two delivery ports P1 and P2, that is, a delivery rate, by controlling a capacity of the capacity control mechanism (swash plate) 2a, which is also referred to a tilt angle of the swash plate. The pump control unit 21 is equipped with a first torque control piston 22a into which a delivery pressure of the delivery port P1 is introduced, a second torque control piston 22b into which a delivery pressure of the delivery port P2 is introduced, and a spring 23 for setting a maximum absorbable torque of the hydraulic pump 2. When an average delivery pressure (P1p+P2p)/2 of the delivery ports P1, P2 of the hydraulic pump 2 has exceeded a predetermined pressure Pa dictated by a biasing force of the spring 23, the pump control unit 21 controls the swash plate 2a of the hydraulic pump 2 to reduce the capacity with increases in the average delivery pressure. The first and second torque control pistons 22a and 22b constitute a torque control device 22 that controls the capacity of the swash plate of the hydraulic pump 2 in such a manner that an absorbable torque of the hydraulic pump 2 does not exceed the maximum absorbable torque that has been set via the spring 23.

Figure 2 is a torque control diagram of the torque control device 22 (the first and second torque control pistons 22a and 22b) of pump 2. Capacity "q" is plotted along a vertical axis in the torque control diagram. The diagram will be a horsepower control diagram if the delivery rate Q (quantity of flow) is instead plotted along the vertical axis.

With reference to Fig. 2, the torque control device 22 remains inactive when the average delivery pressure of the delivery ports P1, P2 is equal to or less than Pa, wherein the average delivery pressure is indicated in Fig. 2 by (P1p+P2p)/2. In this case, a flow control mechanism not shown enables the capacity of the swash plate 2a of the hydraulic pump 2 to be raised to a maximum capacity "qmax" of the hydraulic pump 2 according to a particular amount of manipulation of a control lever (i.e., the flow rate required) without limitation by the torque control device 22.

Once the average delivery pressure of the delivery ports P1, P2 has exceeded Pa, the torque control device 22 operates to conduct limiting control on the tilt angle (capacity) of the hydraulic pump 2 in order to decrease along characteristics line TP1, TP2 as the average delivery pressure rises.

Characteristics line TP1, TP2, geared to the maximum absorbable torque that has been set with the use of the spring 23, is configured to approximate a constant absorbable-torque curve (a hyperbolic curve). The hydraulic pump 2 is thus controlled such that as the average delivery pressure increases, the maximum capacity of the hydraulic pump 2 is reduced along characteristics line TP1, TP2 to maintain the actual absorbable torque at a level equal to or less than the maximum absorbable torque set with the use of the spring 23.

Characteristics line TP1, TP2 in this case is set in such a manner that the maximum absorbable torque of the hydraulic pump 2 is smaller than an output torque TELA of the engine 1 of the present invention, and the torque control device 22 controls the hydraulic pump 2 such that the absorbable torque does not exceed TELA. Accordingly, the absorbable torque of the hydraulic pump 2 stays below the output torque TELA of the engine 1 to prevent the engine from stalling even when an actuator relating to the delivery port P1 of the hydraulic pump 2 and an actuator relating to the delivery port P2 are driven at the same time. In Fig. 2 TELB denotes an output torque of an engine mounted on a conventional seven-ton class hydraulic excavator (TELB will be described later herein).

Figure 3 is an external view of the seven-ton class hydraulic excavator.

The hydraulic excavator in Fig. 3 includes an upper swing structure 300, a lower track structure 301, and a front working implement 302. The upper swing structure 300 is mounted on the lower track structure 301 so as to be swingable, and the front working implement 302 is coupled to a front end of the upper swing structure 300 by means of a swing post 303 so as to be able to turn vertically and horizontally. The lower track structure 301, equipped with left and right crawler belts 310 (opposite side) and 311, includes a vertically movable soil-pushing blade 305 anterior to a track frame 304. The upper swing structure 300 is furnished with a cabin 300a, inside which are provided with a variety of control means, including control levers 309a and 309b for swinging control, as well as for the front working implement (only one of these levers is shown), control levers/pedals 309c and 309d for traveling control (only one of these controls is shown). The front working implement 302 is a structure including a pin-connected boom 306, arm 307, and bucket 308.

The upper swing structure 300 is driven by the swinging motor 9 to swing above the lower track structure 301; the front working implement 302 turns in a horizontal direction by turning the swing post 303 via a swing cylinder 11 (see Fig. 1); the left and right crawler belts 310, 311 of the lower track structure 301 are driven by the left and right traveling motors 8 and 7 (opposite side), respectively, to be rotatable; and the blade 305 is driven by the blade cylinder 10 to be movable upward and downward. In addition, the boom 306, the arm 307, and the bucket 308 pivotally move upward and downward by extending and retracting the boom cylinder 4, the arm cylinder 5, and the bucket cylinder 6, respectively.

Figure 4 is a diagram that lists general specifications of a conventional seven-ton class hydraulic excavator in a tabular format. The specifications shown in Fig. 4 apply to those of a standard type of excavator equipped with a boom of a single type instead of an offset one, and with a bucket as an attachment. With reference to Fig. 4, 1/2 inch is equal to 1.27 cm, 5/8 inch is equal to 1.59 cm, and 3/8 inch is equal to 0.95 cm.

In this Specification, the standard type of hydraulic excavator is defined as a seven-ton class hydraulic excavator having, as shown in Fig. 4, a vehicle weight of 6,300 to 8,500 kg, a bucket size of 0.28 to 0.33 m3, a maximum excavation radius of 6,300 to 7,700 mm, a maximum excavation depth of 3,800 to 4,700 mm, and a maximum excavation height of 6,700 to 7,800 mm.

As shown in Fig. 4, the seven-ton class hydraulic excavators based on related art are equipped with a 40-55 kW engine.

Figure 5 is a diagram that lists gas emission regulations of various countries for seven-ton class excavators. Specific gas emission regulations are defined for each of engine output power ranges. In the EU, the gas emission regulations relating to engines of an output power range of 37<P<56, where P is engine output power, have been strengthened to a level of Stage 11 IB since January 1,2013. For engines of an output power range of 18<P<37, a gas emission regulations level of Stage IMA remains effective in 2013 onward.

The seven-ton class hydraulic excavators based on the related art are equipped with a 40-55 kW engine. This engine output power falls into the 37<P<56 output power range under the gas emission regulations, and after 2013 these engines have been required to comply with EU gas emission regulations Stage NIB. They, therefore, need to have new peripheral devices for reducing the exhaust gas constituents (NOx and PM) to a lower level than ever before. More specifically, a common rail system (CRS), an exhaust gas recirculation (EGR) valve, and a diesel particulate filter (DPF), for example, should be mounted, which makes the engine structure complex and hence increases costs. That also leads to a higher risk of a failure in any of the engine peripherals such as CRS, EGR, and DPF.

In the present invention, output power of the engine 1 mounted on the hydraulic excavator is configured to be 34.0 to 36.4 kW, lower than current 40-55 kW output power. More particularly, the engine output power is set to be 36 kW in the present embodiment. Under the gas emission regulations shown in Fig. 5 the engine output power of 36 kW is classified into the 18<P<37 output power range, and falls into a level of Stage IIIA, one level looser than Stage NIB, under the EU gas emission regulations. In 2013 onward, therefore, the engine further continues to meet the EU gas emission regulations with existing peripheral devices.

The hydraulic pump 2 has its consumption (absorption) horsepower set to be substantially the same as that of a hydraulic pump usually mounted in a seven-ton class hydraulic excavator based on the related art.

In terms of standards, the conduits 4a, 4b, 5a, 5b extending from the control valve 3 to the boom cylinder 4 and the arm cylinder 5 are configured to have a size slightly larger than the hose sizes shown in Fig. 4 for the conventional seven-ton class hydraulic excavators. More specifically, these conduits are arranged with the specifications shown in Fig. 6, that is, the following specifications are applied: size of the rod-end conduit 4a for the boom cylinder 4: 3/4 inch (1.91 cm); size of the bottom-end conduit 4b for the boom cylinder 4: 3/4 inch (1.91 cm); size of the rod-end conduit 5a for the arm cylinder 5: 3/4 inch (1.91 cm); and size of the bottom-end conduit 5b for the arm cylinder 5: 3/4 inch (1.91 cm).

The conduits 7a, 7b, 8a, 8b extending from the control valve 3 to the left and right traveling motors 7, 8 are also configured to have a size larger than the hose sizes shown in Fig. 4 for the conventional seven-ton class hydraulic excavators. More specifically, these conduits are arranged with the conduit specifications of the present invention shown in Fig. 6, that is, the following specifications are applied: size of the conduits 7a, 7b for the right traveling motor 7: 5/8 inch (1.59 cm); and size of the conduits 8a, 8b for the left traveling motor 8: 5/8 inch (1.59 cm).

The operational effects of the present embodiment (i.e., reasons why the working machine properly operates as a seven-ton class hydraulic excavator although the output power range of the engine is lowered by one level) will now be described below for each of functions 1 and 2 separately. - Function 1 -

As described above, in the present embodiment, the output power range of the engine 1 is configured to 36 kW, which falls under the output power range of EU gas emission regulations Stage IIIA. On the other hand, the consumption (absorption) horsepower of the hydraulic pump 2 is set to be substantially the same as that of the related-art hydraulic excavator of the seven-ton class by use of the spring 23.

The output power range of the engine 1 is set to 36 kW, lower than that of the engine usually mounted on the conventional excavator in the present embodiment. On the contrary, as a consequence that the consumption horsepower in the hydraulic pump 2 is set to substantially the same level as that of the hydraulic pump mounted on the conventional excavator, an excess of the pump consumption horsepower over the engine output horsepower (i.e., a distance between characteristics line TP1, TP2 and engine output torque TELA, TELB) decreases below the level observed in the related art. This status is shown in Fig. 2.

Figure 7 is a graph showing a change in the excess of the pump consumption horsepower over the engine output horsepower between the working machines in the related art (left side) and the present embodiment (the present invention - right side).

With a ratio of the excess pump consumption horsepower over the engine output horsepower to the engine output power being defined as a margin ratio, this margin ratio is generally set as follows for the conventional seven-ton class hydraulic excavator:

Margin ratio: 20-30%

The output power range of the engine 1 is set to 36 kW lower than that of the engine mounted on the conventional excavator in the present embodiment. However, the consumption horsepower in the hydraulic pump 2 is set to substantially the same level as that of the hydraulic pump mounted on the conventional excavator, which results in the margin ratio decreasing and being configured to be in a range of 15-20%.

The load on the hydraulic pump is originally configured not to exceed the margin ratio as a possible decline in the engine output caused by variations in usage environment is taken into consideration.

The assumed operating environment for the general seven-ton class hydraulic excavator in the related art is shown below. (a) Operation at high altitudes (b) Operation in a high-temperature environment (increase in intake air temperature) (c) Assumed use of low-quality fuel (d) Increase in fuel temperature

The environment where factors (a) to (d) listed above might be simultaneously occurring is generally expected at the time the margin ratio is established.

In the present embodiment (the present invention), on the other hand, the operating environment is limited to that of the EU and based on the assumption that a degree of the simultaneous occurrence of above factors (a), (b), and (c) is low (the number of highlands and other areas that reach high temperatures is small in the EU). This assumption enables the setting of the margin ratio lower than the general value. - Function 2 -

In addition, the setting of the margin ratio is based on the absorption of an excessive hydraulic pump load caused by a delay in dynamic response of the hydraulic pump.

Figure 8 is a diagram that shows changes with time in the engine output horsepower and pump consumption horsepower of the working machine in the related art. Figure 9 is a diagram that shows changes with time in the engine output horsepower and pump consumption horsepower of the working machine in the present embodiment (the present invention).

As described above, the hydraulic pump 2 includes the torque control device 22 that controls the absorbable torque of the pump to stay equal to or less than the maximum absorbable torque set with the use of the spring 23. The horsepower control function is added to the hydraulic pump 2 in order to keep the consumption (absorption) horsepower below the set level (the consumption horsepower corresponding to characteristics line TP1, TP2 shown in Fig. 9) via the torque control device 22. If the absorbable torque of the hydraulic pump 2 is expressed as τ and the consumption horsepower of the hydraulic pump 2 (the pump consumption (absorption) horsepower) as Jp, a relation of the two can be expressed as follows: τ = (TT/2)*P*q JpOCT P: The delivery pressure of the hydraulic pump 2 q: The capacity of the pump 2

If the delivery pressure of the hydraulic pump 2 rises steeply, a delay in the response of the control function of the torque control device 22 may keep the capacity of the hydraulic pump 2 from decreasing. It is thus assumed that the horsepower absorbed by the hydraulic pump 2 may exceed the pump consumption (absorption) horsepower that has been set. The steep rise in the delivery pressure of the hydraulic pump 2 is prone to arise particularly from a change in the excavating load pressure of the boom cylinder 4 or arm cylinder 5. One of the reasons is that the boom cylinder 4 and the arm cylinder 5 are actuators requiring a higher flow rate of the hydraulic fluid compared with those of other hydraulic actuators. The other one is that both cylinders are driven by the merging of the streams of the fluid delivered from the two delivery ports, P1, P2, of the hydraulic pump 2. Additionally, the flow rates required for the traveling motors 7, 8 are relatively lower than those of the boom cylinder 4 and arm cylinder 5. For this reason, the motors may instantaneously become high in load pressure at the time of the start of travel and are likely in this case to cause the delivery pressure of the hydraulic pump 2 to rise steeply although the motors are each driven by the hydraulic fluid from one of the delivery ports P1 and P2.

In the present embodiment (the present invention), the margin ratio denoting an allowable transient increase in the pump consumption (absorption) horsepower above the control value of the engine 1 output horsepower is lower than in the related art.

As described above, however, the sizes of the conduits 4a, 4b, 5a, 5b, 6a, 6b, 7a, 7b, 8a, 8b for the boom cylinder 4, the arm cylinder 5, and the traveling motors 7, 8, are set to be larger than the general conduit sizes in the present embodiment (the present invention).

The hydraulic hose (conduit) sizes are provided for by the relevant standards. When lines having one size thicker than in the related art are set as so done for the circuits of the boom cylinder 4 and the arm cylinder 5, a cross sectional area of a line becomes approximately twice as large. Energy loss due to a loss of line pressure is therefore reduced to 1/2, or 50%, of that suffered in the related art. That is to say, if the cross sectional area of the line is expressed as A and any loss of energy due to a loss of line pressure is expressed as Jh, a relation between "h" and other related factors can be represented as follows: Q=c*A*AP (a viscous flow of fluid in the line is presumed)

JhocQ*AP=1/(c*A)*Q2 Q: The quantity of fluid flowing through the line ΔΡ: Line pressure loss c: A parameter of the viscous flow A composition ratio of the lines lying in the circuits of the boom cylinder 4 and arm cylinder 5 from the hydraulic pump 2 to the drain tank 12 is 20-30%, which means that the energy loss Jh is reduced by 10% (50% * 20%) of all loss. As shown in Fig. 9, the decrement in margin ratio is corrected and the transient increase in the consumption (absorption) horsepower of the hydraulic pump 2 is suppressed below the output horsepower of the engine 1.

The same as above also applies to the lines lying in the circuits of the traveling motors 7, 8 from the hydraulic pump 2 to the drain tank 12. That is to say, the conduits 7a, 7b, 8a, 8b for the traveling motors 7, 8 are configured to have a size larger than the general conduits.

As a result, an energy loss Jh due to a loss of line pressure is reduced, the decrement in margin ratio is corrected, and the transient increase in the consumption (absorption) horsepower of the hydraulic pump 2 is controlled below the output horsepower of the engine 1.

With the seven-ton class hydraulic excavator according to the present embodiment being configured as described above, the excavator can employ the engine passing the EU gas emission regulations at the Stage MIA level and dispensing with complex peripheral devices (CRS, EGR, DPF). The engine further restrains cost hikes, allowing the excavator to maintain performance comparable to that of the conventional working machine.

While the output power of the engine 1 is set to be 36 kW in the present embodiment, the output power can be larger than 36 kW as long as it is less than 37 kW. Even in this case, the output power of the engine falls into the 18<P<37 output power range under the gas emission regulations, and falls into the level of Stage IMA under the EU gas emission regulations. The engine, therefore, continues to meet the EU gas emission regulations with existing peripheral devices in 2013 onward.

In addition, provided that 34 kW is taken as a lower limit of the engine output power, the decrement in margin ratio is corrected by substantially the same operational effects as achieved when the output power of the engine 1 is set to be 36 kW. The performance comparable to that of the conventional working machine can be maintained consequently.

Furthermore, the sizes of the conduits for the boom cylinder 4, the arm cylinder 5, and the traveling motors 7, 8, are not limited to 3/4 inch (1.91 cm) or 5/8 inch (1.59 cm) and may be other sizes according to which the decrement in margin ratio can be corrected.

The conduits 7a, 7b, 8a, 8b for the traveling motors 7, 8 as well as the conduit size of the boom cylinder 4 and the arm cylinder 5 are configured to have a size larger than the general conduits in the present embodiment. The conduits, however, may be any general-sized ones since the flow rates required for the traveling motors 7, 8 are relatively low relative than those of the boom cylinder 4 and arm cylinder 5 and since these conduits do not suffer significant pressure loss.

Claims

1. A seven-ton class hydraulic excavator with a vehicle weight of 6,300 to 8,500 kg, a bucket size of 0.28 to 0.33 m3, a maximum digging radius of 6,300 to 7,700 mm, a maximum digging depth of 3,800 to 4,700 mm, and a maximum digging height from 6,700 to 7,800 mm, the digging device having an operating environment that is limited to that of the European Union and further comprising: a diesel machine; a hydraulic pump driven by the machine; a plurality of hydraulic actuators, each driven by a hydraulic fluid supplied from the hydraulic pump; and a control valve with a plurality of built-in coils for controlling a flow of the hydraulic fluid supplied by the hydraulic pump to the hydraulic actuators, wherein: the hydraulic actuators comprise a boom cylinder that drives a boom and an arm cylinder that drives an arm; machine output power is set to be 34.0 to 36.4 kW, lower than the current 40 to 55 kW output power of the seven-ton class hydraulic excavator; and a conduit connecting the control valve to the boom cylinder and the arm cylinder is configured to have a size greater than a 1/2 to 5/8 inch (1.27 to 1.59 cm) which is a current pipe size of the hydraulic excavator is of a seven-ton class.

The seven-ton class hydraulic excavator according to claim 1, wherein the line connecting the control valve to the boom cylinder and the arm cylinder has a size of 3/4 inch (1.99 cm).

The seven-ton class hydraulic excavator according to claim 1 or 2, wherein the output power of the machine is 35 kW.

The seven-ton class hydraulic excavator according to claim 1, wherein: the hydraulic actuators further comprise left and right displacement motors for driving left and right. right tracks; and a line connecting the control valve to the left and right displacement motors is configured to have a size greater than 1/2 inch (1.27 cm) which is a current line size of the seven-ton class hydraulic excavator.

The seven-ton class hydraulic excavator according to claim 4, wherein the line connecting the control valve to the left and right displacement motors has a size of 5/8 inch (1.59 cm).

The seven-ton class hydraulic excavator according to one or more of claims 1 to 5, wherein the power consumption of the hydraulic pump is adjusted to be substantially the same as the current power consumption of the seven-ton hydraulic excavator tons of class.