Hydraulic drive system for crawler-mounted construction vehicle

A hydraulic drive system for a crawler-mounted construction vehicle includes: a hydraulic pump (1); a plurality of hydraulic actuators including traveling hydraulic motors (2, 3) driven by the hydraulic pump; a plurality of flow control valve means including first and second flow control valve means (15, 16; 35, 36) for respectively controlling the flow rates of the hydraulic fluid supplied from the pump to motors; pump controlling means (11) for effecting control during the driving of the motors in such a manner that the discharge pressure of the pump becomes higher by a fixed value than the higher one of load pressures of the motors; and first and second pressure balance valve means (17, 18; 37, 38) for respectively controlling the rates of flow through the flow control valve means during the driving of the motors in such a manner that the load pressures of the motors do not affect the rates of flow through the first and second flow control valve means. The first and second pressure balance valve means (17, 18; 37, 38) are respectively provided with valve control means (62, 63; 66, 67 ) which do not cause the associated pressure balance valve means to be actuated until the differential pressure between the load pressure generated in a first motor (2) and the load pressure generated in a second motor (3) reaches a predetermined value (.DELTA.P.sub.0), and causes the pressure balance valve means to be actuated when the differential pressure exceeds the predetermined value.

BACKGROUND OF THE INVENTION 
The present invention relates to a hydraulic drive system for a 
crawler-mounted construction vehicle such as a hydraulic excavator. More 
particularly, the present invention concerns a hydraulic drive system 
having pressure balance valves for effecting load-compensating control of 
the flow rates of hydraulic fluid supplied to a pair of traveling motors 
that are respectively adapted to drive the left and right crawler belts, 
i.e., traveling means for the crawler-mounted construction vehicle. 
As disclosed in U.S. Pat. No. 4,425,759, a known hydraulic drive system for 
a crawler-mounted construction vehicle comprises a hydraulic pump, left 
and right traveling hydraulic motors driven by a hydraulic fluid 
discharged by the hydraulic pump, flow control valves for respectively 
controlling the flow rates of the hydraulic fluid from the hydraulic pump 
to the traveling motors, a pump regulator for effecting control during the 
driving of the left and right traveling motors in such a manner that the 
discharge pressure of the hydraulic pump becomes higher by a fixed value 
than the higher one of load pressures of the left and right traveling 
motors, and pressure balance valves for respectively controlling the rates 
of flow through the flow control valves during the driving of the left and 
right traveling motors in such a manner that the load pressure of the 
traveling motors does not affect the rates of flow through the flow 
control valves. Control of the rates of flow through the flow control 
valves effected by these pressure balance valves is referred to herein as 
load-compensating control. 
In the hydraulic drive system disclosed U.S. Pat. No., 4,425,759, the 
arrangement is such that the pressure balance valves are disposed 
downstream of the associated flow control valves, and the pressure 
downstream of the flow control valves acts in the valve opening direction, 
while the higher one of load pressures of the left and right traveling 
motors acts in the valve closing direction. 
Thus, by providing the pressure balance valves for effecting the 
load-compensating control of the rates of flow through the flow control 
valves for the left and right traveling motors, even if there is any 
differential between the load pressures of the left and right traveling 
motors during traveling, it is possible to effect distribution of fluid 
flow corresponding to the ratio of the valve openings (demand flow rates) 
of the flow control valves for left and right traveling motors. In case of 
steering the crawler belts to change the advancing direction of the 
vehicle, the hydraulic fluid is supplied positively to the higher load 
side traveling motor associated with the outwardly turning crawler belt, 
making it possible to effect an intended change in the course. In 
addition, when the vehicle is made to travel straight, even if the 
resistance to which the left and right crawler belts are subjected 
differs, control is effected in such a manner that the rates flow through 
the flow control valves become equivalent, thereby effecting straight 
traveling. 
With this conventional system, however, since the straight traveling is 
effected by conducting the load-compensating control of the rates of flow 
through the flow control valves by means of the pressure balance valves, 
variations in the performance of hydraulic devices such as the flow 
control valves, the pressure balance valves, etc., that are ascribable to 
fabrication errors affect the straight traveling characteristics. For this 
reason, it has been necessary for the operator to adjust the advancing 
direction while viewing the actual traveling direction. In addition, in 
cases where the strokes of control levers change slightly and the openings 
of the flow control valves is thereby changed, the rates of flow through 
the flow control valves also change, thereby hampering the straight 
traveling characteristics. Consequently, it has been necessary for the 
operator to pay utmost attention to ensure that the strokes of the control 
levers do no change by the slightest degree. Thus, with the conventional 
system, there has been the problem that substantial labor is required of 
the operator in effecting the straight traveling of the vehicle, 
increasing the operator's fatigue. 
In addition, U.S. Pat. No. 4,535,809 discloses a system which is not 
directly related to a hydraulic drive system for a crawler-mounted 
construction vehicle, but in which a flow control valve is arranged by a 
seat valve assembly which comprises a seat-type main valve, a pilot 
circuit associated with the main valve, a pilot valve disposed in this 
pilot circuit and adapted to control the operation of the main valve, and 
a pressure balance valve disposed in the pilot circuit and adapted to 
effect the load-compensating control of the rate of flow through the pilot 
valve. 
In addition, DE-A3422165 discloses a pressure balance valve which is 
disposed upstream of a flow control valve and which causes the discharge 
pressure of a hydraulic pump and the outlet pressure of the flow control 
valve to act in the valve opening direction and causes the maximum one of 
load pressures of a plurality of actuators and the inlet pressure of the 
flow control valve to act in the valve closing direction. 
SUMMARY OF THE INVENTION 
Accordingly, an object of the present invention is to provide a hydraulic 
drive system for a crawler-mounted construction vehicle which is capable 
of positively effecting a change in the advancing direction of the vehicle 
and of readily effecting straight traveling. 
To this end, according to the present invention, there is provided a 
hydraulic drive system for a crawler-mounted construction vehicle 
comprising: a hydraulic pump; a plurality of hydraulic actuators including 
first and second traveling hydraulic motors driven by a hydraulic fluid 
discharged by the hydraulic pump; a plurality of flow control valve means 
including first and second flow control valve means for respectively 
controlling the flow rates of the hydraulic fluid supplied from the 
hydraulic pump to the first and second traveling hydraulic motors; pump 
controlling means for effecting control during the driving of the first 
and second traveling motors in such a manner that the discharge pressure 
of the hydraulic pump becomes higher by a fixed value than the higher one 
of load pressures of the traveling motors; and first and second pressure 
balance valve means for respectively controlling the rates of flow through 
the flow control valve means during the driving of the first and second 
traveling motors in such a manner that the load pressures of the traveling 
motors do not affect the rates of flow through the first and second flow 
control valve means, wherein the first and second pressure balance valve 
means are respectively provided with valve control means which do not 
cause the associated pressure balance valve means to be actuated until the 
differential pressure between the load pressure of the first traveling 
motor and the load pressure of the second traveling motor reaches a 
predetermined value, and causes the pressure balance valve means to be 
actuated when the differential pressure exceeds the predetermined value. 
The valve controlling means preferably comprise springs for urging the 
associated pressure balance valve means in the opening direction. 
In the hydraulic drive system thus arranged in accordance with the present 
invention, when a vehicle is made to travel straight, even if the 
resistance to which the left and right crawler belts are subjected differs 
and a differential arises between the load pressures of the first and 
second traveling motors, the pressure balance valves are not actuated 
insofar as that differential pressure is not more than the aforementioned 
predetermined value, with the result that the first and second traveling 
motors are set in the same condition as that in which the motors are 
connected in parallel. Consequently, in the same way as a general 
hydraulic circuit in which the left and right traveling motors are 
connected in parallel, the flow rates of the hydraulic fluid supplied to 
the first and second traveling motors are forcedly made to be equivalent 
by the forces possessed by the crawler belts themselves to maintain the 
straight traveling, thereby effecting straight traveling. In addition, 
since the straight traveling is effected forcedly by the straight 
traveling-maintaining forces of the crawler belts themselves with the 
pressure balance valve thus set in the nonoperative state, the straight 
traveling is not affected by the variations in the performance of 
hydraulic devices such as the flow control valves and pressure balance 
valves due to fabrication errors and the slightest changes in the strokes 
of the control levers. Hence, it is possible to substantially alleviate 
the burden imposed on the operator. 
Furthermore, when changing the advancing direction of the vehicle, if one 
control lever is operated by a greater degree and the flow rate of the 
hydraulic fluid supplied to the traveling motor associated with the 
outwardly turning-side crawler belt is thereby increased, the differential 
pressure between the load pressures of the first and second traveling 
motors becomes the predetermined value or above, so that the pressure 
balance valve means associated with the lower load pressure side traveling 
motor is actuated. As a result, in the same way as the conventional 
system, it is possible to effect the load-compensating control of the 
rates of flow through the flow control valves, thereby allowing the 
vehicle to turn, as desired. 
With respect to the forms of implementation of the first and second 
pressure balance valve means, a number of examples can be cited. For 
instance, as the first and second pressure balance valve means, it is 
possible to adopt pressure balance valves that are disposed downstream of 
the associated flow control valve means, the pressure balance valves 
having the pressure downstream of the flow control valve means acting in 
the valve opening direction, and the higher one of load pressures of the 
first and second traveling motors acting in the valve closing direction, 
and wherein the valve control means comprise springs for urging the 
associated pressure balance valves in the valve opening direction. 
In addition, it is possible to adopt the first and second flow control 
valve means which include seat valve assemblies each having a seat-type 
main valve, a pilot circuit provided for the main valve, and a pilot valve 
disposed in the pilot circuit and adapted to control the operation of the 
main valve, the first and second pressure balance valve means which 
respectively comprise pressure balance valves each disposed in the pilot 
circuit of the seat valve assembly of the associated flow control valve 
means, the pressure balance valves each being actuated in the valve 
opening direction in response to the pressure differential between the 
discharge pressure of the hydraulic pump and the higher pressure side load 
pressure of the first and second traveling motors and being actuated in 
the valve closing direction in response to the differential pressure 
across the pilot valve, and the valve controlling means which comprise 
springs for urging the associated pressure balance valves in the valve 
opening direction. 
Furthermore, it is possible to adopt the first and second pressure balance 
valve means which respectively comprise pressure balance valves disposed 
upstream of the associated flow control valve means, the pressure balance 
valves each having the differential pressure between the discharge 
pressure of the hydraulic pump and the higher one of load pressures of the 
first and second traveling motors acting in the valve opening direction 
and the differential pressure across the associated flow control valve 
means acting in the valve closing direction, and the valve controlling 
means which comprise springs for urging the associated pressure balance 
valves in the valve opening direction.

DESCRIPTION OF THE PREFERRED EMBODIMENTS 
First Embodiment 
Referring now to FIGS. 1 and 2, a description will be given of an 
embodiment of the present invention. 
In FIG. 1, a hydraulic drive system for a crawler-mounted construction 
vehicle in accordance with the present invention comprises the following 
components: a variable displacement hydraulic pump 1 of, for instance, a 
swash plate type; a plurality of hydraulic actuators that are driven by 
the hydraulic fluid from this hydraulic pump 1, and which include 
hydraulic motors respectively adapted to drive the eft and right crawler 
belts provided in the crawler-mounted construction vehicle, i.e., a left 
traveling motor 2 and a right traveling motor 3; common fluid supply lines 
4, 5 connected to the hydraulic pump 1; common fluid return lines 7, 8 
connected to a tank 6; a main circuit 9 for driving the left traveling 
motor 2 and a main circuit 10 for driving the right traveling motor 3, 
both being connected to the fluid supply line 5 and the return line 7; and 
a pump regulator 11 for controlling the discharge rate of the hydraulic 
pump 1 in such a manner that the discharge pressure of the hydraulic pump 
1 becomes higher by a fixed value than the maximum one of load pressures 
of the plurality of actuators including the left and right traveling 
motors 2, 3. Although not shown, included among the plurality of actuators 
in addition to the left and right traveling motors 2, 3 are hydraulic 
cylinders for driving a boom, an arm, and so forth and/or a hydraulic 
motor for driving a swing in the case where a crawler-mounted construction 
vehicle is a hydraulic excavator, main circuits for driving these 
actuators also being connected to the fluid supply line 5 and the return 
line 7. 
The main circuit 9 for the left traveling motor 2 comprises: two main lines 
13, 14 each having one end connected to the common fluid supply line 5 via 
a fluid supply line 12 and the other end connected to one of the two ports 
of the left traveling motor 2; two control spool valves 15, 16 
respectively disposed in the main lines 13, 14 and adapted to control the 
flow rate of the hydraulic fluid supplied from the hydraulic pump 1 to the 
left traveling motor 2; two pressure balance valves 17, 18 respectively 
disposed downstream of the control spool valves 15, 16 and adapted to 
effect the load-compensating control of the flow rate in such a manner 
that the differential pressure across the control spool valves 15, 16 
becomes substantially fixed; check valves disposed in the main lines 13, 
14 further downstream of the pressure balance valves 17, 18 and adapted to 
permit the flow of the hydraulic fluid directed to the traveling motor 2 
alone; main return lines 22, 23 which respectively branch off from the 
main lines 13, 14 downstream of the check valves 19, 20 and which are 
connected to the common return line 7 via a return line 21; and pressure 
restricting valves 24, 25 respectively disposed in the main lines 22, 23 
for setting back pressure and adapted to vary relief pressure in response 
to the pressure within the other main line 14 or 13. The operation of the 
control spool valves 15, 16 is respectively controlled by pilot pressure 
produced by pilot valves 26, 27, which are operated by a control lever 
(not shown) for the left traveling motor. 
The main circuit 10 for the right traveling motor 3 is arranged in a 
similar manner. In other words, the main circuit 10 comprises: a fluid 
supply line 32; two main lines 33, 34; two control spool valves 35, 36; 
two pressure balance valves 37, 38; check valves 39, 40; a return line 41; 
main return lines 42, 43; and pressure restricting valves 44, 45 for 
setting back pressure. The operation of the control spool valves 35, 36 is 
respectively controlled by pilot valves 46, 47 controlled by a control 
lever (not shown) for the right traveling motor. 
The main circuits of other actuators that are not shown are also arranged 
in a similar manner. 
A pair of check valves 50, 51 which function as higher pressure selecting 
valves are connected between the main lines 13, 14 of the main circuit 9, 
while a pair of check valves 52, 53 which similarly function as higher 
pressure selecting valves are connected to the main lines 33, 34 of the 
main line 10. These check valves 50, 51 and 52, 53 are respectively 
connected to a common load line 56 via load lines 54, 55. Furthermore, 
similar load lines of the main circuits of other actuators are also 
connected to the common load line 56. By this arrangement, the highest one 
of load pressures of the plurality of actuators including the left and 
right traveling motors 2, 3 is introduced into the load lines 54-56 so as 
to detect the maximum load pressure. 
The pressure balance valves 17, 18 comprise valve elements 60, 61 in which 
the output pressure of the control spool valves 15, 16 acts in the valve 
opening direction, while the maximum load pressure acts in the valve 
closing direction, and springs 62, 63 adapted to urge the valve elements 
60, 61 in the valve opening direction. Meanwhile, the pressure balance 
valves 37, 38 similarly comprise valve elements 64, 65 in which the output 
pressure of the control spool valves 35, 36 acts in the valve opening 
direction, while the maximum load pressure acts in the valve closing 
direction, and springs 66, 67 adapted to urge the valve elements 64, 65 in 
the valve opening direction. These springs 62, 63 and 66, 67 serve as 
valve controlling means which operate as follows: During the driving of 
the left and right traveling motors 2, 3, i.e., while traveling, the 
springs 62, 63 and 66, 67 do not allow the valve elements 60, 61 and 64, 
65 to operate until the pressure difference between the load pressure of 
the left traveling motor 2 and that of the right traveling motor 3 reaches 
a predetermined value .DELTA.P.sub.0 determined by the strength of the 
springs, thereby maintaining the pressure balance valves 17, 18 and 37, 38 
in the fully open state. When that pressure differential exceeds the 
predetermined value .DELTA.P.sub.0, the springs 62, 63 and 66, 67 allow 
the valve elements 60, 61 or 64, 65 on the lower load pressure side to 
operate in the closing direction, thereby effecting the load-compensating 
control. 
The pump regulator 11 comprises a hydraulic cylinder 70 for driving a swash 
plate of the hydraulic pump 1 so as to change the displacement volume and 
a control valve 71 for adjusting the displacement of the hydraulic 
cylinder 70. A spring 72 and a pressure-receiving chamber 73 into which 
the maximum load pressure is introduced are provided at one end of the 
control valve 71, while a pressure-receiving chamber 74 into which the 
pump discharge pressure is introduced is provided at the other end 
thereof. When the pressure differential between the discharge pressure of 
the hydraulic pump 1 and the maximum load pressure detected in the load 
line 56 becomes smaller than a setting of the spring 72, the control valve 
71 moves leftward as viewed in the drawing so as to increase the rate of 
discharge of the hydraulic pump 1. When the differential pressure becomes 
greater than the set value of the spring 72, the control valve 71 moves 
rightward as viewed in the drawing, decreasing the rate of discharge of 
the hydraulic pump 1. Thus, the pump regulator 11 controls the rate of 
discharge of the pump 1 in such a manner that the discharge pressure of 
the hydraulic pump 1 becomes higher than the maximum load pressure by a 
fixed value determined by the setting of the spring 62. 
A description will now be given of the operation of this embodiment 
arranged as described above. 
First, the operating characteristics of the pressure balance valves 17, 18 
and 37, 38 will be described. 
As described above, the pressure balance valves 17, 18 are provided with 
springs 62, 63 and 66, 67 serving as valve controlling means which operate 
as follows: During the driving of the left and right traveling motors 2, 
3, i.e., while traveling, the springs 62, 63 and 66, 67 do not allow the 
valve elements 60, 61 and 64, 65 to operate until the pressure difference 
between the load pressure of the left traveling motor 2 and that of the 
right traveling motor 3 reaches the predetermined value .DELTA.P.sub.0, 
thereby maintaining the pressure balance valves 17, 18 and 37, 38 in the 
fully open state. When that pressure differential exceeds the 
predetermined value .DELTA.P.sub.0, the springs 62, 63 and 66, 67 allow 
the valve elements 60, 61 or 64, 65 on the lower load pressure side to 
operate in the closing direction, thereby effecting the load-compensating 
control. Accordingly, when the left and right traveling motors 2, 3 are 
driven by, for instance, opening the control spool valves 15, 35 by the 
same degree of opening, the flow rate of the hydraulic fluid passing 
through the control spool valve 15 or 35 associated with the lower load 
pressure side traveling motor 2 or 3 is controlled in response to an 
increase in the differential between the load pressures of the left and 
right traveling motors 2, 3, as shown in FIG. 2. 
More specifically, when the pressure differential .DELTA.P of the left and 
right traveling motors 2, 3 is not more than a predetermined value 
.DELTA.P.sub.0, neither of the pressure balance valves 17, 37 is 
operative, and both pressure balance valves 17, 37 are maintained in the 
fully open state. At this time, if an increase in the pressure 
differential .DELTA.P is attributable to an increase in maximum load 
pressure, the rate of discharge by the hydraulic pump 1 increases through 
control by the pump regulator 11. Accordingly, the flow rate across the 
control spool valves 15, 35 increases together with increase in the 
differential pressure .DELTA.P. When the differential pressure between the 
load pressures of the left and right traveling motors 2, 3 exceeds the 
predetermined value .DELTA.P.sub.0, the valve element 60 or 64 of the 
pressure balance valve 17 or 37 associated with the lower load pressure 
side traveling motor 2 or 3 is operated in the valve closing direction, 
and its opening is restricted. Consequently, this restricts an increase in 
the flow rate across the control spool valve 15 or 35 resulting from the 
fact that the traveling motor 2 or 3 is on the lower load pressure side, 
and control is hence effected in such a manner that the flow rate across 
the control spool valve 15 or 35 becomes fixed. That is, the flow rate 
across the control spool valve 15 or 35 is subjected to load-compensating 
control in such a manner as to coincide with the flow rate corresponding 
to the valve opening. 
A description will now be given of the overall operation of the hydraulic 
drive system. 
When attempting to effect the straight traveling of the vehicle, the 
operator operates both of control levers (not shown) for right and left 
traveling motors, which in turn actuates the pilot valves 26, 46, thereby 
opening the control spool valves 15, 35 to the degree desired, this degree 
being the same for the two spool valves 15, 35. As a result, the hydraulic 
fluid from the hydraulic pump 1 passes through the control spool valves 
15, 35, passes through the pressure balance valves 17, 37 set in the fully 
open state due to the springs 62, 66, and flows into the left and right 
traveling motors 2, 3, thereby rotating the left and right traveling 
motors 2, 3. At this time, if the resistance to which the left and right 
crawler belts are subjected is identical and the load applied to the left 
and right traveling motors 2, 3 is identical, the load pressures generated 
in the left and right traveling motors 2, 3 also becomes identical. Hence, 
this identical load pressure is detected as the maximum load pressure in 
the load line 56. For this reason, the valve elements 60, 64 of the 
pressure balance valves 17, 37 are not actuated, and the pressure balance 
valves 17, 37 remain fully open, so that equivalent amounts of hydraulic 
fluid corresponding to the openings of the control spool valves 15, 35 are 
supplied to the left and right traveling motors 2, 3. Consequently, the 
left and right traveling motors 2, 3 are caused to rotate at the same 
speed by the supply of this hydraulic fluid, thereby effecting the 
straight traveling of the vehicle. The hydraulic fluid that has rotated 
the left and right traveling motors 2, 3 is returned to the tank 6 via the 
pressure restricting valves 25, 45 for setting back pressure. 
It is now assumed that, during straight traveling by the driving of the 
left and right traveling motors 2, 3, the resistance to which the left 
crawler belt is subjected has become greater than that to which the right 
crawler belt is subjected, and the load pressure of the left traveling 
motor 2 has thereby become greater than that of the right traveling motor 
3. In such a case, the differential between the load pressures acts in the 
valve closing direction on the valve element 64 of the pressure balance 
valve 37 associated with the right traveling motor 3, i.e., the lower load 
pressure side actuator. However, in the range in which that differential 
pressure is smaller than the value .DELTA.P.sub.0 set by the spring 66, 
the valve element 64 of the pressure balance valve 37 is not actuated, as 
described before, and the pressure balance valve 37 is hence kept in the 
fully open state. The differential pressure does not act on the valve 
element of the pressure balance valve 17 associated with the left 
traveling motor 2, i.e., the higher load pressure side actuator, so that 
the pressure balance valve 17 naturally remains in the fully open state. 
Accordingly, the left and right traveling motors 2, 3 assume a state in 
which they are not provided with the pressure balance valves 17, 37, and 
are connected in parallel. Hence, in the same way as the case of a general 
hydraulic circuit in which the left and right traveling motors are 
connected in parallel, the flow rates of hydraulic fluid supplied to the 
left and right traveling motors 2, 3 are forcedly made equivalent by the 
forces presented by the crawler belts themselves to maintain the straight 
traveling of the vehicle, thereby effecting straight traveling. 
Also in a case where the load pressure of the right traveling motor 3 has, 
contrary to the above case, become greater than that of the left traveling 
motor 2, the pressure balance valves 17, 37 are similarly maintained in 
the fully open state, and the flow rates of hydraulic fluid supplied to 
the left and right traveling motors 2, 3 are forcedly made equivalent by 
the forces of the crawler belts themselves to maintain straight traveling, 
thereby effecting straight traveling. 
When changing the advancing direction of the vehicle, e.g., when turning to 
the right, the operator operates the control lever for the left traveling 
motor 2 by a greater degree than the control lever for the right traveling 
motor 3, thereby setting the opening of the control spool valve 15 to a 
greater degree than that of the control spool valve 35. Consequently, with 
respect to the flow rates of hydraulic fluid supplied to the left and 
right traveling motors 2, 3 from the hydraulic pump 1 via the control 
spool valves 15, 35 and the pressure balance valves 17, 37, the flow rate 
for the left traveling motor 2 becomes larger than that of the right 
traveling motor 2, so that the vehicle is moved to turn to the right. 
Thus, when the vehicle is moved to turn to the right, the resistance to 
which the left crawler belt is subjected becomes greater than that to 
which the right crawler belt is subjected. Correspondingly, the load 
pressure of the left traveling motor 2 becomes higher than the load 
pressure of the right traveling motor 3, with the result that a relatively 
large pressure differential not less than the value .DELTA.P.sub.0 set by 
the springs 62, 66 of the pressure balance valves 17, 37 occurs between 
the two load pressures. Consequently, the valve element 64 of the pressure 
balance valve 37 associated with the right traveling motor 3, i.e., the 
lower load pressure side actuator is moved in the valve closing direction. 
Hence, the pressure balance valve 37 operates in a region where the flow 
rate Q is fixed as shown in FIG. 2, thereby controlling the differential 
pressure across the control spool valve 35 to a fixed level, i.e., 
effecting the load-compensating control of the rate of flow across the 
control spool valve 35. As a result of the fact that the rate of flow 
across the control spool valve 35 is thus subjected to load-compensating 
control, the preferential supply of the hydraulic fluid to the right 
traveling motor 3 is restricted, and hydraulic fluid of a fixed flow rate 
corresponding to the opening of the control spool valve 15 is supplied to 
the higher load pressure side left traveling motor 2. Consequently, the 
vehicle turns rightward, as instructed by the control lever. 
As described above, in this embodiment, when an attempt is made to effect 
the straight traveling of the vehicle, since the pressure balance valves 
17, 18 and 37, 38 are set in the nonoperative state, i.e., in the fully 
open state, it is possible to make use of the straight 
traveling-maintaining forces which the left and right crawler belts 
possess. On the other hand, when changing the advancing direction of the 
vehicle, the actuation of the pressure balance valves 17, 18 or 37, 38 
associated with the lower load pressure side traveling motor makes it 
possible to effect a change in the advancing direction by performing the 
load-compensating control of the rate of flow across the control spool 
valve 15, 16, 35, 36. 
In addition, fabrication errors are usually present in hydraulic devices 
constituting a hydraulic drive system, including the control spool valves 
15, 16, 35, 36, the pressure balance valves 17, 18 and 37, 38, etc., and, 
generally speaking, there are variations in the performance on the basis 
of those fabrication errors. If straight traveling is performed by the 
load-compensating control of the pressure balance valves 17, 18 and 37, 
38, these variations in the performance appear as the differential between 
the flow rates of hydraulic fluid supplied to the left and right traveling 
motors 2, 3. Hence, there emerges a situation in which the vehicle fails 
to perform straight traveling despite the fact the operator operates the 
left and right traveling control levers by the same degrees. Accordingly, 
the operator must always operate the control levers while monitoring the 
actual advancing direction, so that substantial labor is required of the 
operator. In addition, if the strokes of the control levers undergo any 
slightest change, the openings of the control spool valves 15, 16, 35, 36 
also change. In this case, if straight traveling is performed by the 
load-compensating control of the pressure balance valves 17, 18 and 37, 
38, the flow rates of hydraulic fluid supplied to the left and right 
traveling motors change, hampering the straight traveling. Accordingly, 
the operator must exercise the utmost care not to move the strokes of the 
control levers once set, and in this case as well, a substantial burden is 
imposed on the operator. 
In contrast, in accordance with this embodiment, since the pressure balance 
valves 17, 18 and 37, 38 are set in the nonoperative state, and the rates 
of flow of hydraulic fluid supplied to the left and right traveling motors 
2, 3 are forcedly made equivalent by the straight traveling-maintaining 
forces of the crawler belts themselves so as to effect straight traveling, 
even if there are any variations in the performance of these hydraulic 
devices or any changes in the positions of the control levers, these 
variations and changes do not appear as the differential between the flow 
rates of hydraulic fluid supplied to the left and right traveling motors, 
and do not affect the straight traveling, making it possible to effect 
straight traveling positively. Accordingly, the burden imposed on the 
operator can be mitigated substantially. 
Second Embodiment 
Referring now to FIGS. 3 and 4, a description will be given of a second 
embodiment of the present invention. In the drawings, those components 
that are equivalent to those shown in FIG. 1 are denoted by the same 
reference numerals. This embodiment is an example in which, instead of the 
spool valves, seat valve assemblies are used as flow control valves. 
In FIG. 3, first and second flow control valves 100, 101 are disposed 
between the hydraulic pump 1 and the left and right traveling motors 2, 3, 
respectively. The flow control valves 100, 101 are respectively composed 
of four seat valve assemblies, i.e., first to fourth seat valve assemblies 
102 to 105 and 102A to 105A. 
In the first flow control valve 100, the first seat valve assembly 102 is 
provided in meter-in circuits 160-162 which are main circuits when the 
left traveling motor 2 is rotated rightward. The second seat valve 
assembly 103 is provided in meter-in circuits 163-165 which are main 
circuits when the left traveling motor 2 is rotated leftward. The third 
seat valve assembly 104 is provided between the left traveling motor 2 and 
the second seat valve assembly 103 in meter-out circuits 165-166 which are 
main circuits when the left traveling motor 2 is rotated rightward. The 
fourth seat valve assembly 105 is provided between the left traveling 
motor 2 and the first seat valve assembly 102 in meter-out circuits 
162-167 which are main circuits when the left traveling motor 2 is rotated 
leftward. 
The meter-in circuit 161 disposed between the first seat valve assembly 102 
and the fourth seat valve assembly 105 is provided with a check valve 110 
for preventing the back flow of the hydraulic fluid flowing to the first 
seat valve assembly, while the meter-in circuit 164 disposed between the 
second seat valve assembly 103 and the third seat valve assembly 103 is 
provided with a check valve 111 for preventing the back flow of the 
hydraulic fluid flowing to the second seat valve assembly. In addition, 
load lines 168, 169 are respectively connected to the upstream side of the 
check valve 110 of the meter-in circuit 161 and the upstream side of the 
check valve 111 of the meter-in circuit 164, a common load line 172 being 
connected to the load lines 168, 169 via check valves 170, 171, 
respectively. 
In the second flow control valve 101 as well, the first to fourth seat 
valve assemblies 102A to 105A are arranged in a similar manner, and a load 
line 172A similar to the load line 172 is provided therein. 
The two load lines 172, 172A are connected to each other via another common 
load line 56, and the highest one of load pressures of the plurality of 
actuators including the left and right traveling motors 2, 3 is introduced 
into the load lines 172, 172A, 56 so as to detect the maximum load 
pressure. 
In the first flow control valve 100, the first to fourth seat valve 
assemblies 102-105 have seat valve-type main valves 112-115, pilot 
circuits 116-115 for the main valves, and pilot valves 120-123 arranged in 
the pilot circuits. The first and second seat valve assemblies 102, 103 
further have pressure balance valves 124, 125 for load-compensating 
control arranged upstream of the pilot valves of the pilot circuits. 
A detailed structure of the first seat valve assembly 102 will be described 
with reference to FIG. 4. 
In the first seat valve assembly 102, the seat-type main valve 112 has a 
valve element 132 for opening and closing an inlet 130 and an outlet 131, 
the valve element 132 being provided with a plurality of slits that 
function as a variable restrictor 133 for changing the opening in 
proportion to the position of the valve element 132, i.e., the opening of 
the main valve. A back pressure chamber 134 communicating with the inlet 
130 via the variable restrictor is formed on the side of the valve element 
132 opposite to the side where the outlet 131 is provided. In addition, 
the valve element 132 is provided with a pressure-receiving portion 132A 
for receiving the inlet pressure of the main valve 112, i.e., the 
discharge pressure Ps of the hydraulic pump 1, a pressure-receiving 
portion 132B for receiving the pressure of the back pressure chamber 134, 
i.e., back pressure Pc, and a pressure-receiving portion 132C for 
receiving the output pressure Pa of the main valve 112. 
The pilot circuit 116 is composed of pilot lines 135-137 allowing the back 
pressure chamber 134 to communicate with the output 131 of the main valve 
112. The pilot valve 120 is composed of a valve element 139 driven by a 
pilot piston 138 to provide a variable restrictor for opening and closing 
a passage between the pilot line 136 and the pilot line 137. The pilot 
piston 138 is driven by pilot pressure produced in correspondence with an 
amount of operation of the control lever (not shown). 
The arrangement of the seat valve assembly having the main valve 112 and 
the pilot valve 120 in combination but not including the pressure balance 
valve 124 is known from U.S. Pat. No. 4,535,809. In this known 
arrangement, if the pilot valve 120 is operated, pilot flow of a rate 
corresponding to the opening of the pilot valve 120 takes place in the 
pilot circuit 116, and this causes the main valve 112 to be opened to a 
degree proportional to the pilot flow rate by virtue of the variable 
restrictor 133 and the back pressure chamber 134, so that the hydraulic 
fluid of a main flow rate amplified in proportion to the pilot flow rate 
flows from the inlet 130 to the outlet 131 via the main valve 112. 
The pressure balance valve 124 for load-compensating control has a valve 
element 140 constituting a variable restrictor, a first pressure-receiving 
chamber 141 for urging the valve element 140 in the valve opening 
direction, and second, third, and fourth pressure-receiving chambers 142, 
143, 144 opposed to the first pressure-receiving chamber and adapted to 
urge the valve element 140 in the valve closing direction. The valve 
element 140 is provided with first to fourth pressure-receiving portions 
145-148 respectively corresponding to the first to fourth 
pressure-receiving chambers 141-144. The first pressure-receiving chamber 
141 is communicated with the back pressure chamber 134 of the main valve 
112 via the pilot lines 149, 135; the second pressure-receiving chamber 
142 is communicated with the pilot line 136; the third pressure-receiving 
chamber 143 is communicated with the maximum load line 172 via the pilot 
line 150; and the fourth pressure-receiving chamber 144 is communicated 
with the inlet 130 of the main valve 112 via the pilot line 152. By virtue 
of this arrangement, the pressure of the back pressure chamber, i.e., the 
back pressure Pc, is introduced into the first pressure-receiving portion 
145; the inlet pressure Pz of the pilot valve 120 is introduced into the 
second pressure-receiving portion 146; the maximum load pressure Pamax is 
introduced into the third pressure-receiving portion 147; and the 
discharge pressure Ps of the hydraulic pump 1 is introduced into the 
fourth pressure-receiving portion. 
The pressure-receiving areas of these pressure-receiving portions are 
determined as follows: If it is assumed that the pressure-receiving area 
of the first pressure-receiving portion 145 is ac, the pressure-receiving 
area of the second pressure-receiving portion 146 is az, the 
pressure-receiving area of the third pressure-receiving portion 147 is am, 
and the pressure-receiving area of the fourth pressure-receiving portion 
148 is as, and that the pressure-receiving area of the pressure-receiving 
portion 132A in the valve element 132 of the aforementioned main valve 112 
is As, the pressure-receiving area of the pressure-receiving portion 132B 
is Ac, and a ratio between the two pressure-receiving areas As/Ac=K (K&lt;1), 
the pressure-receiving areas ac, az, am, and as are set in the proportions 
of 1:1-K:K(1-K):K.sup.2. 
In addition, a spring 153 for urging the valve element 140 in the valve 
opening direction is provided in the first pressure-receiving chamber 141. 
A detailed structure of the second seat valve assembly 103 is identical to 
that of the first seat valve assembly 102, and, in FIG. 3, a spring is 
designated at numeral 154 in correspondence with the spring 153 of the 
pressure balance valve 124. 
Detailed structures of the third and fourth seat assemblies 104, 105 are 
identical to that of the first seat valve assembly 102 with the pressure 
balance valve 124 thereof removed. 
The arrangements of the first to fourth seat valve assemblies 102A-105A in 
the second flow control valve 101 are respectively identical to those of 
the seat valve assemblies 102-105 of the first flow control valve 100. In 
FIG. 3, components of the first to fourth seat valve assemblies 102A-105A 
are denoted by adding "A" to the reference numerals of the corresponding 
seat valve assemblies 102-105, as required. 
In the above-described arrangement, the springs 153, 154, 153A, 154A of the 
pressure balance valves 124, 125, 124A, 125A respectively serve as valve 
controlling means for effecting the load-compensating control of the flow 
rates across the main valves 112, 113, 112A, 113A. Specifically, this 
load-compensating control is carried out as follows: During the driving of 
the left and right traveling motors 2, 3, these springs 153, 154, 153A, 
154A do not actuate the pressure balance valves until the differential 
pressure between their load pressures reaches a predetermined value 
.DELTA.P.sub.0 {=S/K(1-K)} or more, thereby maintaining the pressure 
balance valves in the fully open state, but they actuate the pressure 
balance valve associated with the lower load pressure side traveling motor 
when that differential pressure exceeds the predetermined value 
.DELTA.P.sub.0. 
In the same way as the first embodiment, the hydraulic pump 1 is provided 
with a pump regulator 173 for controlling the discharge rate of the 
hydraulic pump 1 in such a manner that the discharge pressure of the 
hydraulic pump 1 becomes higher by a fixed value than the maximum one of 
load pressures of the plurality of actuators including the left and right 
traveling motors 2, 3. 
The pump regulator 173 comprises a hydraulic cylinder 174 for driving the 
swash plate of the hydraulic pump 1 and changing the displacement volume, 
and a control valve 175 for adjusting the displacement of the hydraulic 
cylinder 174. A spring 176 and a pressure-receiving chamber 177 into which 
the maximum load pressure is introduced are provided at one end of the 
control valve 175, while a pressure-receiving chamber 178 into which the 
pump discharge pressure is introduced is provided at the other end 
thereof. The operation of the hydraulic cylinder 174 and the control valve 
175 is basically identical with that of the hydraulic cylinder 70 and the 
control valve 71 in the first embodiment. 
The operation of this embodiment thus constructed will be described 
hereafter. 
First, a description will be given of the first and second valve assemblies 
102, 103 and 102A, 103A in the first and second flow control valves 100, 
101 by citing the first seat valve assembly 102 as a typical example. 
In the first seat valve assembly 102, the combination of the main valve 112 
and the pilot valve 120 is known, as described above, and main flow takes 
place in the main valve 112 at a rate which is amplified in proportion to 
the pilot flow rate formed in the pilot circuit 116 by the operation of 
the pilot valve 120. When the main valve 112 is thus operative, the 
balance of forces acting on the main valve element 132 can be expressed by 
the following formula from the aforementioned relationships of As/Ac=K(K&lt; 
1): 
EQU Pc=KPs+(1-K) Pa (1) 
Meanwhile, if the balance of forces acting on the valve element 143 in the 
pressure balance valve 124 is considered, since, as described above, the 
pressure-receiving area ac of the pressure-receiving portion 145 is 1, the 
pressure-receiving area az of the pressure-receiving portion 146 is 1-K, 
the pressure-receiving area am of the pressure-receiving portion 147 is 
K(1-K), and the pressure-receiving area as of the pressure-receiving 
portion 148 is K.sup.2, if the force of the spring 153 is assumed to be S, 
the following formula holds: 
EQU Pc+S=(1-K) Pz +K (1-K) Pamax+K.sup.2 Ps (2) 
If the differential pressure Pz-Pa between the inlet pressure and outlet 
pressure of the pilot valve 120 is determined from this formula (2) and 
the aforementioned formula (1), we have 
EQU Pz-Pa=K (Ps-Pamax) +S/(1-K)/ (3) 
Accordingly, the pressure balance valve 124 controls the flow rate across 
the pilot valve 120 in such a manner that the differential pressure across 
the pilot valve 120 agrees with the value of the right-hand side of this 
formula (3). 
Here, in the aforementioned formula (3), the first term of the right-hand 
side, Ps-Pamax, is the differential pressure between the maximum load 
pressure and the discharge pressure of the hydraulic pump 1 controlled by 
the pump regulator 173, and remains constant before the discharge rate of 
the hydraulic pump 1 becomes saturated, but decreases in correspondence 
with the degree of saturation after the discharge rate is saturated. In 
addition, this differential pressure is common to all the pressure balance 
valves 124, 125, 124A, 125A. Furthermore, the first and second terms of 
the right-hand side, the ratio of the pressure-receiving area K and the 
spring force S become common to all the pressure balance valves 124, 125, 
124A, 125A if the ratios of the pressure-receiving area K of the main 
valves 112, 113, 112A, 113A and all the springs 153, 154, 153A, 154A are 
designed to be identical. 
Accordingly, even if a differential arises between the load pressures of 
the left and right traveling motors 2, 3 during the driving of the left 
and right traveling motors 2, 3, the pressure balance valves 124, 125, 
124A, 125A basically control the differential pressure across the pilot 
valves 120, 121, 120A, 121A in such a manner as to be maintained at the 
common and same level so as to control the flow rates across the pilot 
valves 120, 121, 120A, 121A to fixed levels, thereby effecting the 
load-compensating control of the flow rates across the main valves 112, 
113, 112A, 113A. 
The above-described functions are the basic functions of the pressure 
balance valves 124, 125, 124A, 125A. 
The differential pressure across the pilot valve 120 becomes maximum when 
the pressure balance valve 124 is in the fully open state, and this 
maximum differential pressure agrees with Pc-Pa, and therefore this 
differential pressure can be determined from the aforementioned formula 
(1) as follows: 
EQU Pz-Pa =Pc-Pa=K (Ps-Pa) (4) 
If a case is considered where the left and right traveling motors 2, 3 are 
driven simultaneously, and the load pressure of the left traveling motor 2 
is higher than that of the right traveling motor 3, the self load pressure 
agrees with the maximum load pressure Pamax in formulae (4), so formula 
(4) can be translated as: 
EQU Pz-Pa=K (Ps-Pamax) (5) 
Accordingly, through a comparison between formulae (3) and (5), the 
pressure balance valve 124 attempts to control the differential pressure 
across the pilot valve 120 into K (Ps-Pamax)+S/(1-K). However, since the 
differential pressure across the pilot valve 120 does not exceed K 
(Ps-Pamax), so that the pressure balance valve 124 is maintained in the 
fully open state. 
In contrast, in a case where the load pressure of the left traveling motor 
2 is lower than that of the right traveling motor 3, through a comparison 
between formulae (3) and (4), while K (Ps-Pamax)+S/(1-K)&gt;K (Ps-Pa), the 
pressure balance valve 124 attempts to control the differential pressure 
across the pilot valve 120 into K (Ps-Pamax)+S/(1-K). At this time as 
well, since the differential pressure across the pilot valve 120 does not 
exceed K (Ps-Pa), the pressure balance valve 124 is maintained in the 
fully open state. Meanwhile, when K (Ps-Pamax)+S/(1-K)&lt;K (Ps-Pa), it 
becomes possible to control the differential pressure across the pilot 
valve 120 into K (Ps-Pamax)+S/(1-K). Hence, the state of the pressure 
balance valve 124 changes from the fully open state to a restricted state, 
and the pressure balance valve 124 thus controls the differential pressure 
of the pilot valve 120 in such a manner as to make said differential 
pressure agree with K (Ps-Pamax)+S/(1-K), which is a value smaller than 
the maximum value K (Ps-Pa). 
That is, with K (Ps-Pamax)+/(1-K)=K (Ps-Pa) serving as a turning point, the 
pressure balance valve 124 is not actuated in a state in which K 
(Ps-Pamax)+S/(1-K)&gt;K (Ps-Pa), and the pressure balance valve 124 is 
actuated only when K (Ps-Pamax)+S/(1-K)&lt;K (Ps-Pa), effecting the 
load-compensating control of the rate of flow across the pilot valve 120. 
The equation K (Ps-Pamax)+/(1-K)=K (Ps-Pa) can be modified as follows: 
EQU Pamax-Pa=S/K (1-K) (6) 
In this formula (6), the left-hand side represents the differential 
pressure between the load pressure Pa of the left traveling motor 2 and 
the maximum load pressure Pamax (load pressure of the right traveling 
motor 3), while the right-hand side represents a fixed value determined by 
the spring force S of the spring 153 and the area ratio K of the main 
valve 112, and can be replaced by a predetermined value .DELTA.P.sub.0. 
The same operation as that of the pressure balance valve 124 holds true of 
the pressure balance valves 125, 124A, 125A of the second seat valve 
assembly 103 of the first flow control valve 100 and the first and second 
seat valve assemblies 102A, 103A of the second flow control valve 101. 
Thus, the pressure balance valves 124, 125 and 124A, 125A have virtually 
the same operating characteristics as those of the pressure balance valves 
17, 18 and 37, 38 in the first embodiment described with reference to FIG. 
2. 
That is, when the left and right traveling motors 2, 3 are driven by 
opening the pilot valves 120, 120A by the same degrees, in a range in 
which the differential pressure .DELTA.P between the load pressures of the 
left and right traveling motors 2, 3 is not more than .DELTA.P.sub.0 {=S K 
(1-K)}, neither of the pressure balance valves 124, 124A is actuated, and 
both pressure balance valves 124, 124A are maintained in the fully open 
state. At this juncture, if it is assumed that an increase in the 
differential pressure .DELTA.P is due to an increase in the maximum load 
pressure, the discharge rate of the hydraulic pump 1 increases by control 
of the pump regulator 11. Accordingly, the flow rates of the hydraulic 
fluid passing through the main valves 112, 112A increase with an increase 
in the differential pressure .DELTA.P. When the differential pressure 
between the load pressures of the left and right traveling motors 2, 3 
exceeds the predetermined value .DELTA.P.sub.0, the valve element 140 of 
the pressure balance valve 124 or 124A associated with the lower load 
pressure side traveling motor 2 or 3 is moved in the valve closing 
direction, thereby restricting the opening. For this reason, an increase 
of the rates of flow through the pilot valve 120 or 120A and the main 
valve 112 or 112A due to the fact that the traveling motor 2 or 3 is the 
lower load pressure side is suppressed, and control is effected in such a 
manner that the rate of flow through the main valve 112 or 112A becomes 
fixed. In other words, the rate of flow through the main valve 112 or 112A 
is subjected to load-compensating control in such a manner as to 
correspond to a flow rate corresponding to the opening of the pilot valve 
120 or 120A. 
As described above, since the pressure balance valves 124, 125, 124A, 125A 
function in the same way as the pressure balance valves of the first 
embodiment, the hydraulic drive system in accordance with this embodiment 
operates in a manner similar to that of the first embodiment, as described 
below. 
In short, when the vehicle is made to travel straight, control levers (not 
shown) for left and right traveling motors are operated to cause the pilot 
valves 120, 120A to open by the same degrees, for instance. Consequently, 
pilot flow of equal flow rates takes place in the pilot circuits 116, 
116A, while flow of equal rates amplified in proportion thereto takes 
place in the main valves 112, 112A as well, thereby supplying the 
hydraulic fluid at equal flow rates to the left and right traveling motors 
2, 3. At this time, when the resistance to which the left and right 
crawler belts are subjected differs, and when there is a difference in the 
load pressure between the left and right traveling motors 2, 3, the 
pressure balance valves 124, 124A are not actuated and remain in the fully 
open state if the differential pressure is less than the predetermined 
value .DELTA.P.sub.0. Hence, the left and right motors 2, 3 remain in the 
same condition as that in which they are connected in parallel. 
Consequently, in the same way as a general hydraulic circuit in which left 
and right traveling motors are connected in parallel, the flow rates of 
the hydraulic fluid supplied to the left and right traveling motors 2, 3 
are forcedly made identical by the straight traveling-maintaining forces 
possessed by the crawler belts themselves, thereby effecting the straight 
traveling. 
In addition, since the straight traveling is forcedly effected by the 
straight traveling-maintaining forces of the crawler belts themselves with 
the pressure balance valves 124, 124A thus set in the nonoperative state, 
even if there are any variations in the performance of the hydraulic 
devices such as the main valves 112, 112A, the pilot valves 120, 120A, the 
pressure balance valves 124, 124A due to the fabrication errors or any 
slightest changes in the strokes of the control levers, insofar as the 
differential pressure between the load pressures is not more than the 
predetermined value .DELTA.P.sub.0, unnecessary load-compensating control 
is not effected and the intended straight traveling is not hampered. 
When the advancing direction of the vehicle is to be changed, for instance, 
when the vehicle is to be turned to the right, one control lever is 
operated by a greater degree than the other control lever, so that, for 
instance, the opening of the pilot valve 120 of the first seat valve 
assembly 102 is set to a greater degree. As a result, the amount of 
hydraulic fluid supplied to the left traveling motor 2 increases, so that 
the differential pressure between the load pressures of the left and right 
traveling motors 2, 3 becomes the predetermined value .DELTA.P.sub.0 or 
more, and the pressure balance valve 124A associated with the lower load 
pressure side right traveling motor 3 is actuated. The differential 
pressure across the pilot valve 120A is thereby controlled to a fixed 
level, so that an increase in the rate of flow through the main valve 112A 
is suppressed, and the hydraulic fluid is supplied to the left and right 
traveling motors 2, 3 at a rate corresponding to the amounts of opening of 
the pilot valves 120, 120A. Hence, the vehicle turns to the right, as 
instructed by the control levers. 
Thus, in this embodiment as well, it is possible to effect the straight 
traveling of the vehicle and a change in its advancing direction. At the 
same time, the straight traveling is effected by making use of the 
straight traveling-maintaining forces possessed by the crawler belts 
themselves with the pressure balance valves 124, 125 and 124A, 125A set in 
the nonoperative state. Accordingly, even if there are variations in the 
performance of the hydraulic devices or any changes in the strokes of the 
control levers, it is readily possible to effect the straight traveling 
with such variations and changes exerting an influence on the straight 
traveling. Hence, it is possible to substantially mitigate the burden 
imposed on the operator when effecting a straight traveling. 
In the above-described second embodiment, as described above, the following 
arrangement is adopted as the meter-in circuit-side seat valve assemblies 
102, 103 and 102A, 103A: The pressure balance valves 124, 125, 124A, 125A 
for load-compensating control are disposed upstream of the pilot valves 
120, 121, 120A, 121A. The pressure balance valve is provided with the 
first pressure-receiving portion 145 for urging the valve in the valve 
opening direction and the second, third and fourth pressure-receiving 
portions 146-148 for urging the valve in the valve closing direction. The 
pilot valve inlet pressure Pz, the maximum load pressure Pamax, and the 
pump discharge pressure Ps are introduced into these pressure-receiving 
portions 145-148. The pressure-receiving areas of these pressure-receiving 
portions are set in the proportions of 1:1-K:K (1-K):K.sup.2. However, on 
June 30, 1988, the assignee of the present invention filed an application 
for patent for an invention concerning a flow control valve comprising 
seat valve assemblies each having a special load compensating function, as 
Japanese Patent Application No. 63-163646. Thus, various modifications of 
the construction and arrangement of the seat valve assembly are possible 
in accordance with the concept of the invention of this earlier 
application. 
For instance, with respect to the pressure balance valve, it suffices if 
the aforementioned formula (3) ultimately holds. In other words, the 
pressure balance valve may take any form or arrangement insofar as the 
pressure balance valve is actuated in the valve opening direction in 
response to the pressure differential between the discharge pressure of 
the hydraulic pump 1 and the higher one of load pressures of the left and 
right traveling motors 2, 3, is actuated in the valve closing direction in 
response to the differential pressure across the pilot valve, and is 
provided with a spring for urging the valve in the valve opening 
direction. 
Third Embodiment 
Referring now to FIG. 5, a description will be given of a third embodiment 
of the present invention. 
In the drawing, those members that are equivalent to the members shown in 
FIG. 1 are denoted by the same reference numerals. This embodiment is an 
example in which a general load compensating valve is used as the pressure 
balance valve. 
In FIG. 5, the hydraulic drive system in accordance with this embodiment 
comprises the following components: main lines 200-202 constituting a main 
circuit for driving the left traveling motor 2; main lines 203-205 
constituting a main circuit for driving the right traveling motor 3; a 
flow control valve 206 disposed in the main lines 200-202 and adapted to 
control the flow rate of the hydraulic fluid supplied from the hydraulic 
pump 1 to the left traveling motor 2; a pressure balance valve 207 
disposed in the main line 200 upstream of the flow control valve 206 and 
adapted to control the rate of flow through the flow control valve 206 in 
such a manner that the differential pressure across the flow control valve 
206 becomes substantially fixed, thereby effecting load-compensating 
control; a check valve 208 disposed between the flow control valve 206 and 
the pressure balance valve 207 and adapted to permit only the flow of the 
hydraulic fluid directed to the traveling motor 2; a flow control valve 
209 disposed in the main lines 203-205 and adapted to control the flow 
rate of the hydraulic fluid supplied from the hydraulic pump 1 to the left 
traveling motor 3; a pressure balance valve 210 disposed in the main line 
203 upstream of the flow control valve 209 and adapted to control the rate 
of flow through the flow control valve 209 in such a manner that the 
differential pressure across the flow control valve 209 becomes 
substantially fixed, thereby effecting load compensating control; and a 
check valve 211 disposed between the flow control valve 209 and the 
pressure balance valve 210 and adapted to permit only the flow of the 
hydraulic fluid directed to the traveling motor 2. The operation of the 
flow control valves 206, 209 is controlled by pilot pressure produced by 
pilot valves respectively operated by control levers (not shown). 
In addition, the discharge rate of the hydraulic pump 1 is controlled by 
the pump regulator 11 in such a manner that the discharge pressure of the 
hydraulic pump 1 becomes higher by a fixed value than the maximum load 
pressure of the plurality of actuators including the left and right 
traveling motors 2, 3. 
The flow control valves 206, 209 have load ports 212, 213 respectively 
communicated with the main lines 201, 202 and 204, 205 in the left and 
right switched positions illustrated in the drawing for detecting the load 
pressure of the left and right traveling motors 2, 3. The load ports 212, 
213 are respectively connected to load lines 214, 215. The load lines 214, 
215 are further connected to a shuttle valve 216 which functions as a 
higher pressure selecting valve, the shuttle valve 216 being connected to 
the common load line 56 via a load line 217. Similar load lines of the 
main circuits of other actuators (not shown) are connected to the common 
load line 56. By virtue of the above-described arrangement, the highest 
one of load pressures of the plurality of actuators including the left and 
right traveling motors 2, 3 is introduced into the load lines 56, 217. 
The pressure balance valve 207 has two pressure-receiving portions 218, 219 
for urging a valve element within the pressure balance valve 207 in the 
valve opening direction, and two pressure-receiving portions 220, 221 for 
urging the same in the valve closing direction. The discharge pressure of 
the hydraulic pump 1 is introduced from a hydraulic line 223 into the 
pressure-receiving portion 218. The load pressure of the meter-in circuit 
of the left traveling motor 2, i.e., the outlet pressure of the flow 
control valve 206, is introduced into the pressure-receiving portion 219 
via a hydraulic line 224. The maximum load pressure is introduced into the 
pressure-receiving portion 220 via a hydraulic line 225. The inlet 
pressure of the flow control valve 206 in the meter-in circuit is 
introduced into the pressure-receiving portion 221 via a hydraulic lie 
226. All the pressure-receiving areas of the pressure-receiving portions 
218-221 are set to be identical. 
The pressure balance valve 210 is also arranged in a similar manner. In 
other words, the pressure balance valve 210 has two pressure-receiving 
portions 227, 228 for urging the valve in the valve opening direction and 
two pressure-receiving portions 229, 230 for urging the valve in the valve 
closing direction. The discharge pressure of the hydraulic pump 1 is 
introduced into the pressure-receiving portion 227 via a hydraulic line 
232. The load pressure of the meter-in circuit of the right traveling 
motor 3, i.e., the outlet pressure of the flow control valve 209 in the 
meter-in circuit, is introduced into the pressure-receiving portion 228 
via a hydraulic line 233. The maximum load pressure is introduced into the 
pressure-receiving portion 229 via a hydraulic line 234. The inlet 
pressure of the flow control valve 209 in the meter-in circuit is 
introduced into the pressure-receiving portion 230 via a hydraulic line 
235. 
The pressure balance valves 207, 210 are further provided with springs 22, 
231 for urging valve elements disposed within the pressure balance valves 
in the valve closing direction. In the same way as the springs 62, 63 and 
66, 67 of the first embodiment, these springs 222, 231 function as valve 
control means as follows: During the driving of the left and right 
traveling motors 2, 3, i.e., while traveling, these springs 222, 231 do 
not actuate the pressure balance valves 207, 210 until the differential 
pressure between the load pressure of the left traveling motor 2 and the 
load pressure of the right traveling motor 3 reaches a predetermined value 
.DELTA.P.sub.0 set by the strength of the spring, thereby maintaining the 
same in the fully open state, and the springs 222, 231 actuate the 
pressure balance valve associated with the lower load pressure side 
traveling motor when that differential pressure exceeds the predetermined 
value .DELTA.P.sub.0, thereby effecting load-compensating control. 
By virtue of the above-described arrangement, during the driving of the 
left and right traveling motors 2, 3, when the opening of the flow control 
valves 206, 209 is fixed, the pressure balance valve 207 or 210 associated 
with the lower load pressure side traveling motor 2 or 3 controls the rate 
of flow through the associated flow control valve 206 or 209, as 
illustrate in FIG. 2. 
In short, in the range in which the differential pressure .DELTA.P between 
the load pressures of the left and right traveling motors 2, 3 is less 
than the predetermined value .DELTA.P.sub.0, the forces acting in the 
valve opening direction in the pressure balance valves 207, 210 are 
greater than the forces acting in the valve closing direction, so that 
neither of the pressure balance valves 207, 210 is actuated, and the 
pressure balance valves 207, 210 are maintained in the fully open state. 
At this juncture, if it is assumed that an increase in the differential 
pressure .DELTA.P is due to an increase in the maximum load pressure, the 
discharge rate of the hydraulic pump 1 increases by control of the pump 
regulator 11. Accordingly, the rates of flow through the flow control 
valves 206, 209 increase with an increase in the differential pressure 
.DELTA.P. When the differential pressure between the load pressures of the 
left and right traveling motors 2, 3 exceeds the predetermined value 
.DELTA.P.sub.0, with respect to the pressure balance valve 207 or 210 
associated with the lower load pressure side traveling motor 2 or 3, the 
force acting in the valve closing direction becomes greater than that 
acting in the valve opening direction, thereby moving the valve in the 
valve closing direction and, hence, restricting the valve opening. For 
this reason, an increase in the rates of flow through the flow control 
valves 206, 209 due to the fact that the traveling motor 2 or 3 is on the 
lower load pressure side is suppressed, and control is thus effected in 
such a manner that the rate of flow through the flow control valve 206 or 
209 becomes fixed. In short, the rates of flow through the flow control 
valves 206, 209 are subjected to load-compensating control in such a 
manner as to agree with flow rates corresponding to the valve openings. 
As described above, since the pressure balance valves 207, 210 function in 
the same way as the pressure balance valves of the first embodiment, the 
hydraulic drive system in accordance with this embodiment also operate in 
the same way as that of the first embodiment, as described below. 
In other words, when the vehicle is made to travel straight, the control 
levers (not shown) for left and right traveling motors are operated to 
open the flow control valves 206, 209, for instance, by the same degrees 
in the switched positions on the right-hand side as viewed in the drawing. 
As a result, the hydraulic fluid of the same flow rates flows to the flow 
control valves 206, 209, supplying the hydraulic fluid to the left and 
right traveling motors 2, 3 at the same flow rates. At this time, when the 
resistance to which the left and right crawler belts are subjected 
differs, and there is a differential in the load pressure between the left 
and right traveling motors 2, 3, the pressure balance valves 207, 210 are 
not actuated and remain in the fully open state if that differential 
pressure is less than the aforementioned predetermined value 
.DELTA.P.sub.0. Hence, the left and right traveling motors are in the same 
condition as that in which the motors 2, 3 are connected in parallel. For 
this reason, in the same way as a general hydraulic circuit in which the 
left and right traveling motors are connected in parallel, the flow rates 
of the hydraulic fluid supplied to the left and right traveling motors 2, 
3 are forcedly made equivalent by the straight traveling-maintaining 
forces possessed by the crawler belts themselves, thereby effecting the 
straight traveling. In addition, since the straight traveling is effected 
forcedly by the straight traveling-maintaining forces of the crawler belts 
themselves with the pressure balance valves 207, 209 set in the 
nonoperative state, even if there are any variations in the performance of 
the hydraulic devices such as the flow control valves 206, 209, the 
pressure balance valves 207, 210, etc., due to fabrication errors or any 
slightest changes in the strokes of the control levers, unnecessary 
compensating control is not effected insofar as the differential pressure 
between the load pressures is not more than the predetermined value 
.DELTA.P.sub.0, thereby making it possible to effect the intended straight 
traveling. 
When changing the advancing direction of the vehicle, e.g., when turning to 
the right, one control lever is operated by a greater degree than the 
other control lever, so that, for instance, the opening of the flow 
control valve 206 is set to a greater degree. Consequently, the amount of 
the hydraulic fluid supplied to the left traveling motor 2 increases, so 
that the differential pressure between the load pressures of the left and 
right traveling motors 2, 3 becomes the aforementioned predetermined value 
.DELTA.P.sub.0 or above, and the pressure balance valve 210 associated 
with the lower load pressure side right traveling motor 3 is actuated. As 
a result, the differential pressure across the flow control valve 209 is 
controlled to a fixed level, so that an increase in the rate of flow 
therethrough is suppressed, and the hydraulic fluid is supplied to the 
left and right traveling motors 2, 3 at flow rates corresponding to the 
openings of the flow control valves 206, 209. Consequently, the vehicle 
turns to the right, as instructed by the control levers. 
As described above, in this embodiment as well, it is possible effect the 
straight traveling of the vehicle and a change in its advancing direction. 
At the same time, since straight traveling is effected by the straight 
traveling-maintaining forces possessed by the crawler belts themselves 
with the pressure balance valves 207, 210 set in the nonoperative state, 
even if there are variations in the performance of the hydraulic devices 
or any changes in the strokes of the control levers, it is readily 
possible to effect the straight traveling without such variations and 
changes affecting the straight traveling. Hence, it is possible to 
substantially alleviate the burden imposed on the operator when effecting 
straight traveling. 
In the above-described third embodiment, the arrangement is such that the 
pump discharge pressure and the maximum load pressure are directly 
introduced hydraulically into the pressure balance valves 207, 210. 
However, an arrangement may be alternatively provided such that, instead 
of directly introducing the pump discharge pressure and the maximum load 
pressure hydraulically, a common differential pressure gauge for detecting 
the differential pressure between the two pressures is provided, and a 
detected signal of this differential pressure gauge may be imparted 
electrically to the pressure balance valves or after being converted into 
a hydraulic signal. 
As described above, although three embodiments of the present invention 
have been given, the present invention is not restricted to these 
embodiments, and various modifications are possible without departing from 
the spirit of the present invention. For instance, in the foregoing 
embodiments, springs are used as valve controlling means which do not 
cause the associated pressure balance valves to be actuated until the 
differential pressure between the load pressures of the left and right 
traveling motors reaches a predetermined value, and causes the pressure 
balance valve to be actuated when that differential pressure exceeds the 
predetermined value. However, an arrangement may be alternatively provided 
such that a fixed level of hydraulic force is imparted instead of using 
the springs.