Patent Publication Number: US-6668549-B2

Title: Hydrostatic vehicle driving system applicable to a working vehicle

Description:
BACKGROUND OF THE INVENTION 
     1. Field of the Invention 
     The present invention relates to a hydrostatic vehicle driving system applicable to a working vehicle like that having a mower unit, namely, a lawn tractor, including a hydrostatic transmission preferably incorporated in a transaxle apparatus, wherein the speed reduction ratio exerted by the driving system is automatically changed in correspondence to the condition of load on an engine. 
     2. Related Art 
     Conventionally, it is well-known that a working vehicle like a lawn tractor having a mower unit is equipped with a belt type stepless transmission having a split pulley (hereinafter, “a belt transmission”) for avoiding overload on an engine of the vehicle. Generally, such a belt transmission is interposed between the output shaft of the engine and the input shaft of the transaxle apparatus incorporating a mechanical transmission like a multi-speed gear type transmission. The belt transmission automatically changes its speed reduction ratio according to the change of output rotary speed of the engine. In this regard, the speed reduction ratio of the belt transmission is smaller when the engine drives fast, and greater when the engine drives slowly. 
     This system is advantageous in that, while the engine drives fast, the smaller speed reduction ratio is established for efficient exertion of the capacity of the engine, and that, while the engine drives slow, the engine can be automatically prevented from overload because of the greater speed reduction. However, the tandem arrangement of the belt transmission and the transaxle apparatus interferes with minimization and cost-saving of the vehicle. 
     On the other hand, for being applied to a working vehicle such as a lawn tractor, there is a well-known conventional transaxle apparatus (an integral hydrostatic transmission, namely, an “IHT”) which incorporates a hydrostatic stepless transmission (hereinafter, an “HST”). The speed reduction ratio exerted by the HST, which corresponds to the ratio of capacity of a hydraulic motor to that of a hydraulic pump in the HST, is conventionally changed by a driver&#39;s manual operation of a movable swash plate of the variable capacity hydraulic pump. However, the conventional IHT used in such a working vehicle is not provided with a device automatically changing the speed reduction ratio of its HST in correspondence to output of or load on an engine. 
     Incidentally, even if the speed reduction ratio of such an HST in IHT for a working vehicle can be automatically changed correspondingly to load on an engine, there are still some cases such that the automatic changing of the speed reduction ratio is not desired. Particularly, if the working vehicle is a lawn tractor, it is preferable that, during its mowing work, the speed reduction ratio is not automatically changed, but manually changed only by a driver&#39;s will for preventing the lawn from being mowed unevenly. 
     BRIEF SUMMARY OF THE INVENTION 
     A main object of the present invention is to provide a hydrostatic vehicle driving system between an engine and an axle, used in a working vehicle like a lawn tractor, including an HST which comprises a hydraulic pump and a hydraulic motor fluidly connected with each other through a hydraulic circuit, wherein the speed reduction ratio exerted by the driving system can be automatically steplessly changed in correspondence to load on the engine without the use of a conventional belt transmission having a split pulley for minimization and cost-saving of the vehicle. 
     To achieve the object, the hydraulic motor is made to be a variable displacement hydraulic motor, whose capacity is controlled by a motor capacity control system in correspondence to the condition of load on the engine. 
     Consequently, the speed reduction ratio of the HST is automatically optimally increased for avoiding overload on the engine, and reduced for effectively transmitting power of the engine to the axle. The present vehicle driving system with an HST can be improved in its minimization and cost-saving because there is no use of the belt transmission as mentioned above. 
     Preferably, the hydraulic pump, the hydraulic motor and the axle are contained in a housing, thereby constituting a compactly integrated transaxle apparatus. 
     The motor capacity control system comprises load-detection means detecting hydraulic pressure in the hydraulic circuit corresponding to the load on the engine, a hydraulic actuator for changing the capacity of the hydraulic motor, and actuator-control means controlling the hydraulic actuator according to the hydraulic pressure detected by the load-detection means. 
     Accordingly, the capacity of the hydraulic motor can be changed without an expensive electric sensor or actuator, but with hydraulic oil utilized as it is used in the HST or the transaxle apparatus, thereby enabling the motor capacity control system to be provided simply and at low cost. 
     If there are some cases that the vehicle having the above-mentioned present hydrostatic vehicle driving system is desired to cruise at a constant speed, the automatically controlled capacity of the hydraulic motor is preferred to be fixed. 
     Then, the present hydrostatic vehicle driving system comprises a manual mode selection member which is provided on a working vehicle so as to be switched between a first mode position and a second mode position. When the mode selection member is located at the first mode position, the capacity of the hydraulic motor is fixed, and when the mode selection member is located at the second mode position, the capacity of the hydraulic motor can be varied by the motor capacity control system. 
     If the hydraulic motor is of an axial piston type, the hydraulic motor is provided with a movable motor swash plate having a contact surface abutting against a piston of the hydraulic motor. The motor swash plate is moved from a first angle to a second angle larger than the first angle accordingly to increase the load detected by the load detection means. The first angle and the second angle are respectively formed between the contact surface of the motor swash plate and the phantom plane perpendicular to a rotary axis of the hydraulic motor. 
     When the motor swash plate is located at the first angle, the capacity of the hydraulic motor is smaller so as to establish the smaller speed reduction ratio of the HST. When the motor swash plate is located at the second angle, the capacity of the hydraulic motor is greater so as to establish the greater speed reduction ratio of the HST. 
     Particularly, if the vehicle is a lawn tractor, it is preferable that the first mode position corresponds to its traveling during lawn-mowing and the second mode position corresponds to its regular traveling on a road or the like. If the variation of capacity of the hydraulic pump is out of consideration, then when the vehicle travels for mowing, the mode selection member is located at the first mode position so as to fix the speed reduction ratio, thereby enabling the vehicle to cruise at a constant speed so as to prevent the lawn from being mowed unevenly. When the vehicle travels on a road or the like out of mowing-work, the mode selection member is located at the second mode position so as to change the speed reduction ratio in correspondence to the detected load on the engine, thereby enabling an effective speed control and prevention of overload on the engine. 
     The hydraulic pump is a variable displacement hydraulic pump, and a manual speed control member is provided on the working vehicle for changing capacity of the hydraulic pump. 
     The capacity of the hydraulic pump is controlled by operation of the speed control member whether the mode selection member is located at the first mode position or the second mode position. In other words, when the mode selection member is located at the first mode position, the speed reduction ratio is not automatically controlled by the motor capacity control means, but manually changed only by operation of the speed control means. When the mode selection member is located at the second mode position, the speed reduction ratio is automatically controlled by the motor capacity control means in addition to its manual change by operation of the speed control means. 
     If the hydraulic pump is an axial piston type variable displacement hydraulic pump, the manual speed control member is operated so as to move a movable pump swash plate of the hydraulic pump. 
     The speed control member may be selectively connected to a carburetor of the engine so that, when the mode selection member is located at the second mode position, the speed control member is operated so as to control both capacity of the hydraulic pump and output rotary speed of the engine. Especially, the speed control member may be selectively connected to a throttle member of the carburetor so that, when the mode selection member is located at the second mode position, the speed control member is operated so as to move both the pump swash plate and the throttle member. 
     Therefore, on the above-mentioned assumption that the vehicle is a lawn tractor, during its regular traveling out of mowing work, only the speed control member is manipulated without manipulation of an accelerator member provided on the vehicle, thereby enabling the engine and transmission in the vehicle to be controlled with one hand. If the vehicle is to ascend a slope or start while the mode selection member is located at the second mode position, the engine is desirably accelerated in association with the shift of the transmission to highspeed by such an easy operation. 
     For an alternative mechanism for capacity control of the variable displacement hydraulic pump in association with output control of the engine, an actuator for changing capacity of the hydraulic pump may be provided in addition to the manual speed control member so as to be controlled according to the output of the engine, and a selection means is provided for selecting one of the manual speed control member and the actuator so as to change the capacity of the hydraulic pump. This selection means may be provided in association with the mode selection member so that, while the selection means selects the speed control member, capacity of the hydraulic motor is fixed, and that, while the selection means selects the actuator, capacity of the hydraulic motor can be changed. 
     In this regard, the above-mentioned hydraulic circuit for fluidly connecting the hydraulic pump and the hydraulic motor with each other serves as a first hydraulic circuit for driving the hydraulic motor, and additionally, a second hydraulic circuit is extended from a discharge port of a charge pump driven by the engine for supplying the first hydraulic circuit with fluid. An orifice is provided on the way of the second hydraulic circuit, so that the actuator is driven according to difference of hydraulic pressure in the second hydraulic circuit between upstream and downstream of the orifice. 
     Consequently, when the selection means selects the actuator, capacity of the hydraulic pump is varied according to operation of a manual accelerator member such as a pedal or a lever for adjusting the throttle of a carburetor of the engine, thereby enabling both engine controlling and transmission shifting with one hand (foot). 
     Other and further objects, features and advantages of the present invention will appear more fully from the following description. 
    
    
     BRIEF DESCRIPTION OF THE DRAWINGS/FIGURES 
     FIG. 1 is a diagram of an entire hydrostatic vehicle driving system having a motor capacity control system  200  according to the present invention; 
     FIG. 2 is a plan view of a transaxle apparatus  1  incorporating an HST  8 , from which an upper housing part  9   a  except for a part thereof is removed, being applicable to the hydrostatic vehicle driving system as shown in FIG. 1; 
     FIG. 3 is a sectional view taken along the lines III-UI of FIG. 2; 
     FIG. 4 is a sectional view taken along the lines IV—IV of FIG. 2; 
     FIG. 5 is an enlarged fragmentary plan view of the same transaxle apparatus  1  from which upper housing part  9   a  except for a part thereof is removed, showing a principal portion of a hydraulic motor  21  therein, wherein a slant angle of a movable motor swash plate  23  is set to the minimum; 
     FIG. 6 is an enlarged fragmentary plan view of the same, wherein the slant angle of the motor swash plate  23  is set to the maximum; 
     FIG. 7 is a diagram of a principal portion of a hydrostatic vehicle driving system having a motor capacity control system  200 ′ according to the present invention; 
     FIG. 8 is a plan view of transaxle apparatus  1  incorporating HST  8 , from which upper housing part  9   a  except for a part thereof is removed, being applicable to the hydrostatic vehicle driving system as shown in FIG. 7; 
     FIG. 9 is a sectional view taken along the lines IX—IX of FIG. 8; 
     FIG. 10 is a sectional view taken along the lines X—X of FIG. 8, wherein a slant angle of motor swash plate  23  is set to the minimum; 
     FIG. 11 is a sectional view taken along the lines X—X of FIG. 8, wherein the slant angle of motor swash plate  23  is set to the maximum; 
     FIG. 12 is a sectional view taken along the lines XI—XI of FIG. 9; 
     FIG. 13 is a diagram of a principal portion of a hydrostatic vehicle driving system having a motor capacity control system  200 ′ a  as a modification of motor capacity control system  200 ′ according to the present invention; 
     FIG. 14 is a diagram of a principal portion of a hydrostatic vehicle driving system having a motor capacity control system  200 ′ b  as a modification of motor capacity control system  200 ′ according to the present invention; 
     FIG. 15 is a diagram of an entire hydrostatic vehicle driving system having a motor capacity control system  200 ′ c  as a modification of motor capacity control system  200 ′ according to the present invention, wherein an accelerator system  100  is linked through a link mechanism  300  with a speed control pedal  27  for changing capacity of hydraulic pump  11 ; 
     FIG. 16 is a diagram of a principal portion of a first embodiment of the hydrostatic vehicle driving system as shown in FIG. 15 from which link mechanism  300  for connecting accelerator system  100  with speed control pedal  27  is removed, wherein speed control pedal  27  is neutral and a mode selection lever  36  is located at a work mode position m 1 ; 
     FIG. 17 is a diagram of the same, wherein a fore pedal portion  27   a  of speed control pedal  27  is pressed and mode selection lever  36  is located at work mode position m 1 ; 
     FIG. 18 is a diagram of the same, wherein fore pedal portion  27   a  of speed control pedal  27  is pressed, mode selection lever  36  is located at regular traveling mode position m 2 , and load applied on an engine  2  is less than the considerable; 
     FIG. 19 is a diagram of the same, wherein fore pedal portion  27   a  of speed control pedal  27  is pressed, mode selection lever  36  is located at a regular traveling mode position m 2 , and considerable load applied on engine  2  is detected; 
     FIG. 20 is a diagram of a principal portion of a second embodiment of the hydrostatic vehicle driving system as shown in FIG. 15, showing that link mechanism  300  for connecting accelerator system  100  with speed control pedal  27  is provided so as to perform both engine-control and transmission-shift by operation of only speed control pedal  27 , wherein speed control pedal  27  is neutral and mode selection lever  36  is located at work mode position m 1 ; 
     FIG. 21 is a diagram of the same, wherein fore pedal portion  27   a  of speed control pedal  27  is pressed and mode selection lever  36  is located at work mode position m 1 ; 
     FIG. 22 is a diagram of the same, wherein fore pedal portion  27   a  of speed control pedal  27  is pressed, mode selection lever  36  is located at regular traveling mode position m 2 , and load applied on engine  2  is less than the considerable; 
     FIG. 23 is a diagram of the same, wherein fore pedal portion  27   a  of speed control pedal  27  is pressed, mode selection lever  36  is located at regular traveling mode position m 2 , and considerable load applied on engine  2  is detected; 
     FIG. 24 is a sectional view taken along the lines IX—IX of FIG. 8, when FIG. 8 serves as a plan view of transaxle apparatus  1  from which upper housing part  9   a  except for a part thereof is removed, being applicable to the hydrostatic vehicle driving system as shown in FIGS. 15 to  23 ; 
     FIG. 25 is a sectional view taken along the lines X—X of FIG. 8, when FIG. 8 serves as a plan view of transaxle apparatus  1  from which an upper housing part  9   a  except for a part thereof is removed, being applicable to the hydrostatic vehicle driving system as shown in FIGS. 15 to  23 , wherein the slant angle of motor swash plate  23  is set to the minimum; 
     FIG. 26 is a similar sectional view, wherein the slant angle of motor swash plate  23  is set to the maximum; 
     FIG. 27 is a diagram of an entire hydrostatic vehicle driving system having motor capacity control system  200 ′ c  according to the present invention, wherein an accelerator system  100 ′ is provided and an automatic speed control system  160  can be selectively connected to a pump swash plate  13  through selection means  150  so that the capacity of hydraulic pump  11  can be controlled according to the throttle operation of carburetor  130  of engine  2 ; 
     FIG. 28 is a diagram of a principal portion of the hydrostatic vehicle driving system as shown in FIG. 27, wherein mode selection lever  36  is located at work mode position m 1 , speed control pedal  27  is neutral and the load on engine  2  is less than the considerable; 
     FIG. 29 is a diagram of the same, wherein mode selection lever  36  is located at regular traveling mode position m 2 , engine  2  does not drive, and the load on engine  2  is less than the considerable; 
     FIG. 30 is a diagram of the same, wherein mode selection lever  36  is located at regular traveling mode position m 2 , engine  2  drives at some speed, and load applied on an engine  2  is less than the considerable; 
     FIG. 31 is a diagram of the same, wherein mode selection lever  36  is located at work mode position m 1 , fore pedal portion  27   a  of speed control pedal  27  is pressed, and the load on engine  2  is less than the considerable; 
     FIG. 32 is a diagram of the same, wherein mode selection lever  36  is located at regular traveling mode position m 2 , engine  2  does not drive, and considerable load applied on engine  2  is detected; 
     FIG. 33 is a sectional view taken along the lines IX—IX of FIG. 8, when FIG. 8 serves as a plan view of transaxle apparatus  1  from which upper housing part  9   a  except for apart thereof is removed, being applicable to the hydrostatic vehicle driving system as shown in FIGS. 27 to  34 ; and 
     FIG. 34 is a perspective view of selection means  150  with mode selection lever  36 . 
    
    
     DETAILED DESCRIPTION OF THE INVENTION 
     Basically, a hydrostatic vehicle driving system according to the present invention includes an HST having a variable displacement hydraulic motor whose capacity is controlled with hydraulic pressure in a closed fluid circuit of the HST in correspondence to engine load. This driving system is particularly applicable to a working vehicle such as that having a mower unit, namely, a lawn tractor. 
     A basic style of the hydrostatic vehicle driving system of the present invention will be described in accordance with FIG.  1 . 
     Power from an engine  2  mounted on a working vehicle is transmitted to a pair of left and right axles  50 L and  5 OR through an HST  8 , a gear train  30  and a differential  40 . Generally, axles  50 L and  5 OR are driven for driving rear wheels of the working vehicle. HST  8  includes a variable displacement hydraulic pump  11  and a variable displacement hydraulic motor  21  fluidly connected with each other through a pair of first and second hydraulic oil passages  111  and  112 . 
     Between engine  2  and hydraulic pump  11  is interposed a regular belt transmission without a split pulley. In this regard, an output pulley  4  is fixed on an output shaft  3  of engine  2 . An input pulley  6  is fixed on an input shaft (a pump shaft)  12  of hydraulic pump  11 . A belt  5  is interposed between both pulleys  4  and  6 . 
     Hydraulic pump  11 , which is of an axial piston type, has a movable pump swash plate  13 . A speed control pedal  18  for speed changing and a forward/backward travel selection lever  28  for switching the traveling direction of the vehicle between forward and backward are provided in the vicinity of a driver&#39;s seat of the working vehicle. 
     In this embodiment, speed control pedal  18  is also linked with a carburetor of engine  2  so that the opening of carburetor is increased in proportion to the pressing of speed control pedal  18 . Consequently, the vehicle is accelerated by increasing compound outputs of both engine  2  and HST  8 . 
     Forward/backward travel selection lever  28  is linked with a fitting linkage interposed between pump swash plate  13  of hydraulic pump  11  and speed control pedal  18 . When forward/backward travel selection lever  28  is switched, the linkage between pump swash plate  13  and speed control pedal  18  is switched so as to reverse the slanting direction of pump swash plate  13 , whereby the direction of oil discharged from hydraulic pump  11  to hydraulic motor  21  is switched, thereby switching the traveling direction of the working vehicle between forward and backward. 
     Alternatively, as shown in FIG. 15, speed control pedal  18  and forward/backward travel selection lever  28  may be combined into a double speed control pedal  27  like a seesaw. As shown in FIG.  16  and others, double speed control pedal  27  is pivoted at its intermediate portion, and provided with a pair of pedal portions, which are a fore pedal portion  27   a  and a rear pedal portion  27   b , arranged in opposite with respect to the intermediate pivot portion. A swing arm  27   c  is extended from the intermediate pivot portion of pedal  27 , and linked to a control arm  61  for rotating motor swash plate  23  through a link rod  51 . Depending upon whether fore pedal portion  27   a  or rear pedal portion  27   b  is pressed, control arm  61  is rotated to one of opposite sides from its neutral position so as to rotate motor swash plate  23  into either its forward traveling range or backward traveling range. 
     In each of later described hydrostatic vehicle driving systems, speed control pedal  27  may be exchanged with speed control pedal  18  and forward/backward switching lever  28  in correspondence to the variation of structure of the working vehicle. 
     FIG. 1 simply illustrates that speed control pedal  18  is linked with an engine carburetor so as to be used as an accelerator pedal. This interlocking of pump swash plate  13  of hydraulic pump  11  with the engine carburetor may be performed in both cases of forward and backward traveling of the working vehicle. Alternatively, if the working vehicle is not required to travel backward at high speed, this interlocking may be established only when the working vehicle travels forward and released when the working vehicle travels backward. Of course, hydraulic pump  11  may be out of interlocking with the engine carburetor so as to control the output of HST  8  independently to the output of engine  2 . 
     An accelerator system  100  shown in FIGS. 15 to  23  will be described. A carburetor  130  of engine  2  is provided with a throttle arm  134 . A manual accelerator lever  20  is fixedly provided with a swing arm  20   a . Swing arm  20   a  is linked with throttle arm  134  through a wire  64  so that the angle of throttle arm  134  is adjusted by rotating accelerator lever  20 , thereby controlling the output rotary speed of engine  2 . 
     Accelerator lever  20  is biased toward its neutral (idling) position by a return spring  52 . Furthermore, accelerator lever  20  is provided with friction lock means  45  so as to be frictionally held at the operated position against return spring  52  after it is released from an operator&#39;s force. 
     Accelerator system  100  shown in FIGS. 15 to  19  is free from the operation of speed control pedal  27 . FIGS. 20 to  23  illustrates accelerator system  100  which can selectively interlock with speed control pedal  27  through a link mechanism  300  (also shown in FIG.  15 ). In link mechanism  300 , speed control pedal  27  and friction lock means  45  interlock with a later-discussed mode selection lever  36 . 
     Alternatively, FIGS. 27 to  32  illustrates another accelerator system  100 ′. In accelerator system  100 ′, accelerator lever  20  is integrally provided with a plate portion  20   a . An arcuate slot  20   b  centering on the pivot of accelerator lever  20  is bored through plate portion  20   a . A slide pin  20   c  is slidably engaged in slot  20   b  and connected to throttle arm  134  through a wire  64 . Slide pin  20   c  is initially held at one end of slot  20   b . This accelerator lever  20  is also provided with friction lock means  45 . 
     When accelerator lever  20  is manually rotated to an optimal position from its idling position, slide pin  20   c , while being held at the one end of slot  20   b , is moved together with accelerator lever  20  so as to rotate throttle arm  134 . After the operation of accelerator lever  20  is finished and released from a driver&#39;s force, accelerator lever  20  is held at the optimal position by friction lock means  45 . 
     Furthermore, a momentary accelerator pedal  46  is disposed in the vicinity of a driver&#39;s seat. Momentary accelerator pedal  46  is integrally provided with a swing arm  46  which is connected with slide pin  20   c  through a wire  65 . Momentary accelerator pedal  46  is biased to its neutral (idling) position by a return spring  52 . 
     When momentary accelerator pedal  46  is pressed against return spring  52 , slide pin  20   c  slides from the one end to the other end in slot  20   b  while accelerator lever  20  is held by friction lock means  45 , thereby rotating throttle arm  134  from the position determined by lever  20  as long as pedal  46  is pressed. Afterward, when momentary accelerator pedal  46  is released, pedal  46  returns to the idling position by spring  52  and throttle arm  134  returns to the position determined by accelerator lever  20  fixed by friction lock means  45 . 
     The operation of each of accelerator systems  100  and  101 ′ in association with speed control pedal  27 , mode selection lever  36  and the like will be detailed in later descriptions of each corresponding hydrostatic vehicle driving system for a working vehicle. 
     Now, a transaxle apparatus  1  as an integrated hydrostatic transmission (an IHT) which is applicable to the hydrostatic vehicle driving system used in a working vehicle as shown in FIG. 1 will be described in accordance with FIGS. 1 to  6 . 
     As shown in FIGS. 2 to  6 , transaxle apparatus  1  comprises a housing  9  constituted by an upper housing part  9   a  and a lower housing part  9   b  which are joined with each other through their surrounding horizontal flat joint surfaces. A bearing portion for a motor shaft  22  is formed by upper and lower housing parts  9   a  and  9   b  on their joint surfaces. Axles  50 L and  50 R are rotatably supported through a bearing portion formed by upper housing part  9   a  above its joint surface. In housing  9  is disposed differential  40  through which axles  50 L and  50 R are differentially connected at their inside ends with each other. Axles  50 L and  50 R project laterally outwardly from left and right outer ends of housing  9 , respectively. 
     As shown in FIG. 2, the inner space of housing  9  is divided into a first chamber R 1  and a second chamber R 2  through a partition wall  9   i  formed integrally-with housing  9 . Moreover, partition wall  9   i  is integrally formed of upper housing part  9   a  and lower housing part  9   b  joined with each other. In first chamber R 1  is disposed HST  8 . In second chamber R 2  are disposed differential  40 , gear train  30  which is interposed between motor shaft  22  of HST  8  and differential  40 , and axles  50 L and  50 R. 
     Partition wall  9   i  extends laterally along axle  50 R and is bent so as to extend longitudinally in perpendicular to axle  50 R, thereby making first chamber R 1  substantially rectangular and forming second chamber R 2  in a substantially L-like shape disposed along two adjacent edges of first chamber R 1 . Therefore, HST  8  and axle  50 R are juxtaposed forwardly and rearwardly through a lateral extending portion of partition wall  9   i , and HST  8  and gear train  30  are juxtaposed rightwardly and leftwardly through a longitudinal extending portion of partition wall  9   i . Differential  40  is disposed in the corner portion of L-like shaped second chamber R 2  serving as a cross point of axles  50 L and  50 R and gear train  30 . 
     First chamber R 1  and second chamber R 2  are filled with oil in common, thereby serving as oil sumps. Partition wall  9   i  is bored at its optimal portion by a hole in which an oil filter  81  is disposed. For example, oil filter  81  may be provided in the lateral extending portion of partition wall  9   i  between HST  8  and axle  50 R as shown in FIGS. 2 and 3. Oil filter  81  allows oil to flow therethrough between first and second chambers R 1  and R 2 . First chamber R 1  is supplied therein with oil cleaned through oil filter  81  from second chamber R 2 , while the oil undergoing obstacles such as iron powder generated from rubbing gears in second chamber R 2 . Thus, the oil used as lubricating oil for gears and bearings of gear train  30 , differential  40  and the like can be also used as fine operating oil for HST  8 . 
     Housing  9  is provided at the ceiling of its upper housing part  9   a  with an oil replenishing port (not shown) in communication with first chamber R 1 . An external oil reservoir (not shown) disposed outside housing  9  is connected with the oil replenishing port directly or through a piping (not shown) like a rubber hose. While the oil used as operating oil for HST  8  is heated by the activation of HST  8  so as to expand the whole of oil in first chamber R 1  (and second chamber R 2 ), the excessively increased oil flows into the oil reservoir, thereby adjusting the volume of oil in housing  9 . 
     Referring to gear train  30 , as shown in FIGS. 1,  2  and  4 , in second chamber R 2 , an output gear  31  and a brake disc  32  are fixed onto motor shaft  22  of HST  8 , and a brake device  33  is disposed in the vicinity of brake disc  32  so as to apply braking force onto brake disc  32  for braking motor shaft  22 . 
     A counter shaft  39  is rotatably disposed parallel between motor shaft  22  and differential  40  in second chamber R 2 . Counter shaft  39  is formed on its periphery with a speed reduction gear  38 . A speed reduction gear  37  which is diametrically larger than speed reduction gear  38  is disposed around counter shaft  39 . Speed reduction gear  37  is formed at its inner periphery with teeth which engages with speed reduction gear  38 , thereby being fixed to counter shaft  39 . Speed reduction gear  37  engages with output gear  31 . Speed reduction gear  38  engages with an input gear  41  of differential  40 . 
     Referring to differential  40 , each of coaxial left and right axles  50 L and  50 R is fixedly provided on its inward end portions with a side bevel gear  44 . Axles  50 L and  50 R are further extended toward each other from respective side bevel gears  44  thereon and slidably rotatably inserted into a central through-hole of input gear  41 . 
     Input gear  41  is bored through between left and right side surfaces thereof with a pair of pinion holes  48  disposed symmetrically with respect to the central through-hole. A pinion shaft  49  and a bevel pinion  43  are disposed in each pinion hole  48 . Each bevel pinion  43  is provided on pinion shaft  49  through a friction member  56  so as to apply a certain frictional braking force onto rotating bevel pinion  43 , thereby designating differential  40  as a limited-slip differential. 
     Differential  40  is provided with a differential-locking system. In this regard, a lock member  47  is slidably provided around one axle  50  (in this embodiment, right axle  50 R). Pawls  47   a  provided on lock member  47  are engaged in respective holes  42  of input gear  41 . Lock member  47  is slidable along axle  50 R while it engages with input gear  41  through pawls  47   a  and holes  42 . One side bevel gear  44  (fixed on right axle  50 R) is formed with recesses  44   a . By sliding lock member  47  along axle  50 R, lock member  47  is engaged/disengaged with/from side bevel gear  44  through recesses  44   a . Accordingly, differential  40  is locked by an operator so as to make axles  50 L and  50 R rotatable integrally with each other. 
     HST  8  in transaxle apparatus  1  shown in FIGS. 1 to  6  has such a structure as follows: 
     Center section  10  is removably attached to housing  9  in first chamber R 1 . Center section  10  is provided at its forward half portion with a vertical surface disposed perpendicularly to axles  50 L and  50 R. The vertical surface serves as a motor mounting surface  10   m  to which hydraulic motor  21  is mounted. Center section  10  is provided at its rearward half portion with a horizontal surface serving as a pump mounting surface  10   p  to which hydraulic pump  11  is mounted. In hydraulic pump  11 , a cylinder block  14  is slidably mounted onto pump mounting surface  10   p  so as to be rotatable centering its vertical axis relative to center section  10 . A pump shaft  12  is vertically disposed through a center of pump mounting surface  10   p  and rotatably supported by center section  10 . Pump shaft  12  is axially disposed through cylinder block  14 . Pump shaft  12  and cylinder block  14  are locked together. Cylinder block  14  are provided therein with a plurality of cylinder holes disposed in parallel to pump shaft  12  and on the periphery of pump shaft  12 . Pistons  15  with biasing springs are reciprocally movably inserted into the cylinder holes, respectively. 
     Pump shaft  12  projects upwardly from the top of upper housing part  9   a  for serving as an input shaft. Input pulley  6  and a cooling fan  7  are fixed onto the upwardly projecting portion of pump shaft  12 . As mentioned above, as shown in FIG.  6  and other drawings showing each hydrostatic vehicle driving system discussed later, output pulley  4  is fixed onto output shaft  3  of vehicle engine  2 . Belt  5  is interposed between output pulley  4  and input pulley  6  so as to transmit power from engine  2  to hydraulic pump  11 . 
     Movable pump swash plate  13  is disposed in housing  9  between the ceiling of upper housing part  9   a  and cylinder block  14  so as to abut against heads of pistons  15 . Pump swash plate  13  is tilted so as to incline its surface abutting against the heads of pistons  15  at an optimal angle from the horizontal plane perpendicular to the rotary axis of cylinder block  14 , thereby varying the direction and amount of oil discharged from hydraulic pump  11  to hydraulic motor  21  through later-discussed hydraulic circuit consisted by a pair of hydraulic oil passages  111  and  112  within center section  10 . 
     Pump swash plate  13  made as a trunnion type is downwardly curved at its both ends on which coaxial trunnion shafts  60  and  60   a  are laterally provided respectively. Trunnion shaft  60   a  is journalled in partition wall  9   i . Trunnion shaft  60  is journalled through a cover  9   c  attached to upper housing part  9   a  and extended outwardly for serving as a speed control shaft, as shown in FIG.  2 . Speed control arm  61  is fixed onto the outward projecting portion of trunnion shaft  60 . 
     For constituting the linkage among pump swash plate  13 , speed control pedal  18  and forward/backward traveling selection lever  28  as shown in FIG. 1, speed control arm  61  is linked with both speed control pedal  18  and forward/backward traveling selection lever  28 . 
     By pressing speed control pedal  18 , speed control arm  61  together with trunnion shafts  60  and  60   a  are rotated in a longitudinal direction of the working vehicle so as to rotate pump swash plate  13  slantwise around trunnion shafts  60  and  60   a . Thus, the amount of oil discharged from hydraulic pump  11  to hydraulic motor  21  is varied so as to change the output rotary speed of hydraulic motor  21 , thereby changing the traveling speed of the working vehicle in traveling either forward or backward. 
     By switching lever  28  between a forward traveling position and a backward traveling position when pump swash plate  13  is located at a neutral position, the rotational direction of pump swash plate  13  corresponding to the depth of pressed speed control pedal  18  is changed oppositely with respect to the surface of pump swash plate  13  perpendicular to pistons  15 . Thus, when lever  28  is located at the forward traveling position, pump swash plate  13  is rotated in its rotational range for forward traveling from the neutral position according to the depth of pressed speed control pedal  18 . When lever  28  is located at the backward traveling position, the opposite happens. 
     As shown in FIGS. 2 and 5, pump swash plate  13  is formed integrally with a neutral arm  13   a , from which a pin  67  projects toward cover  9   c . An eccentric shaft  66  is disposed through cover  9   c  and projects inwardly and outwardly from cover  9   c . The inward projecting portion of eccentric shaft  66  is disposed eccentrically to the axis of the portion thereof supported in cover  9   c . In housing  9 , a neutral spring  69  is coiled around trunnion shaft  60  and extended parallel at its both end portions so as to sandwich pin  67  and the inward projecting portion of eccentric shaft  66 . 
     When speed control arm  61  is rotated from its neutral position, pin  67  is integrally rotated so as to forcedly push one of the end portions of neutral spring  69  further away from the other end portion which is held at its initial position by eccentric shaft  66 . When rotated speed control arm  61  is released from the operating force which has been applied thereon, the end portion of neutral spring  69  which has been pushed by pin  67  returns to its initial position by its biasing force so as to push pin  67  to its initial position, thereby automatically returning pump swash plate  13  to its neutral position. 
     A neutral adjusting nut  66   a  is provided on the threaded outward projecting portion of eccentric shaft  66 . When eccentric shaft  66  is rotated in relative to nut  66   a , the inward projecting portion of eccentric shaft  66  disposed between both the end portions of neutral spring  69  is revolved centering the axis of the portion of eccentric shaft  66  disposed in cover  9   c  so as to change the initial position of both the end portions of neutral spring  69  together with pin  67 , thereby adjusting the neutral position of pump swash plate  13 . 
     In hydraulic motor  21 , a cylinder block  24  is slidably mounted onto motor mounting surface  10   m  so as to be rotatable relative to center section  10  centering its horizontal axis disposed in parallel to axles  50 L and  50 R. Motor shaft  22  is axially disposed through cylinder block  24 , thereby being disposed laterally in parallel to axles  50 L and  50 R. Motor shaft  22  and cylinder block  24  are locked together. Cylinder block  24  are provided therein with a plurality of cylinder holes disposed in parallel to motor shaft  22  and on the periphery of motor shaft  22 . Pistons  25  with biasing springs are reciprocally movably inserted into the cylinder holes, respectively. 
     Motor shaft  22  is horizontally disposed between upper and lower housing parts  9   a  and  9   b  as mentioned above. Motor shaft  22  is inserted at one end thereof into center section  10  through the center of pump mounting surface  10   p  and rotatably supported by center section  10 . Motor shaft  22  is extended through a movable motor swash plate  23  from cylinder block  24  oppositely to center section  10  so as to project at the other end thereof into second chamber R 2 . Motor shaft  22  is journalled at its intermediate portion through a bearing  29  in partition wall  9   i . Bearing  29  is sandwiched between the portions of upper and lower housing parts  9   a  and  9   b  which are formed into partition wall  9   i . Bearing  29  is provided with an oil seal so as to prevent oil from mutually flowing between chambers R 1  and R 2  therethrough. 
     Motor swash plate  23  can be rotated between a minimum slant angle position and a maximum slant angle position. When motor swash plate  23  is located at the minimum slant angle position, the surface of motor swash plate  23  abutting against the heads of pistons  25  is slanted at a minimum angle A 1  from a plane P which is perpendicular to the rotary axis of hydraulic motor  21  (motor shaft  22  and pistons  25 ), whereby the amount of oil discharged from hydraulic motor  21  is the minimum. At this time, if the slant position of pump swash plate  13  is out of consideration, a speed reduction ratio established by HST  8  is the minimum. 
     When motor swash plate  23  is located at the maximum slant angle position, the surface of motor swash plate  23  abutting against the heads of pistons  25  is slanted at a maximum angle A 2  from plane P, whereby the amount of oil discharged from hydraulic motor  21  is the maximum. At this time, if the slant position of pump swash plate  13  is out of consideration, the speed reduction ratio of HST  8  is the maximum. 
     The slant position of motor swash plate  23  is changed by a hydraulic actuator in correspondence to the load applied on engine  2 . The hydraulic control system for motor swash plate  23  is distinctive in each transaxle apparatus  1  among the present embodiment thereof shown in FIGS. 1 to  6  and other embodiments described later. Each of the different hydraulic control systems of motor swash plate  23 , namely, a motor capacity control system  200  or the like, will be discussed later. 
     First and second hydraulic oil passages  111  and  112  formed in center section  10  serve as a closed oil circuit for fluidly connecting hydraulic pump  11  and hydraulic motor  21  with each other. In center section  10 , a pair of kidney ports (not shown) are open at pump mounting surface  10   p  and a pair of first and second kidney ports  95  and  96  are open at motor mounting surface  10   m . In center section  10  is bored along its longitudinal direction with an upper horizontal oil hole  91  and a lower horizontal oil hole  92  in parallel. A slant oil hole  90  is formed slantwise in center section  10  and connected with upper horizontal oil hole  91 . The first kidney port on pump mounting surface top is connected with first kidney port  95  on motor mounting surface  10   m  through oil holes  91  and  90 . The second kidney port on pump mounting surface  10   p  is connected with second kidney port  96  on motor mounting surface  10   m  through oil hole  92 . 
     A series of oil holes  91  and  90  constitute first hydraulic oil passage  111 . Oil hole  92  serves as second hydraulic oil passage  112 . Consequently, hydraulic oil is circulated between hydraulic pump  11  and hydraulic motor  21  through the pair of first and second hydraulic oil passages  111  and  112 . 
     When first hydraulic oil passage  111  is hydraulically higher-pressured and second hydraulic oil passage  112  is hydraulically depressed, hydraulic motor  21  is rotated in one direction for driving axles  50 L and  50 R forward. When second hydraulic oil passage  112  is hydraulically higher-pressured and first hydraulic oil passage  111  is hydraulically depressed, hydraulic motor  21  is rotated in the other opposite direction for driving axles  50 L and  50 R backwardly. Such a difference of hydraulic pressure between first and second hydraulic oil passages  111  and  112  is mainly established by setting the position of pump swash plate  13  of hydraulic pump  11  with the above-mentioned operation means for speed changing and determining the traveling direction of the working vehicle, like speed control pedal  18  and forward/backward traveling selection lever  28  and so on. 
     The rotational force of hydraulic motor  21  is transmitted into axles  50 L and  50 R through gear train  30  and differential  40 . 
     HST  8  is provided with a structure for compensation of hydraulic oil in first and second hydraulic oil passages  111  and  112 . Center section  10  is bored with a vertical charge oil passage  93  crossing upper and lower horizontal oil holes  91  and  92 . A check valve  26  is disposed in a crossing point between each of oil holes  91  and  92  and charge oil passage  93  for prevention of backflow of oil to each of hydraulic oil passages  111  and  112 . 
     An open bottom end of charge oil passage  93  at the bottom surface of center section  10  serves as a charge oil port. The charge oil port is connected with a discharge port of a charge pump  16  which is disposed below center section  10 . Charge pump  16  is a usual trochoid pump. A charge pump casing  16   a  is attached onto a bottom surface of center section  10 . An inner rotor and an outer rotor are disposed in charge pump casing  16   a . A suction filter  17  is disposed in housing  9  and attached to an inlet opening of the suction port of charge pump casing  16   a  as shown in FIG.  3 . Pump shaft  12  of hydraulic pump  11  is extended downwardly through center section  10  so as to drive the inner and outer rotors of charge pump  16 . Therefore, pump shaft  12  also serves as a drive shaft for charge pump  16 . 
     As shown in FIG. 1, a relief valve  76  is provided in charge pump casing  16   a  for regulating the charge pressure effecting in the discharge port of charge pump casing  16   a.    
     As shown in FIGS. 1 and 3, a check valve  19  for prevention of freewheel is provided in charge pump casing  16   a  so as to be interposed between the discharge port of charge pump  16  and charge oil passage  93 . 
     When the working vehicle is stationary on a slope, hydraulic motor  21  receives a driving force from axles  50 L and  50 R and functions as a pump so as to reduce the hydraulic oil in either hydraulic oil passage  111  or  112 . However, check valve  19  is opened by negative pressure of either hydraulic oil passages  111  or  112  so as to absorb the oil in housing  9 , thereby preventing the reduction of hydraulic oil. 
     When the working vehicle is towed, oil is bypassed between hydraulic oil passages  111  and  112 . 
     In this regard, as shown in FIGS. 2,  4  and  5 , a camshaft  77  is disposed vertically and rotatably supported by upper housing part  9   a . Camshaft  77  projects upwardly from housing  9  so as to be fixedly provided thereon with a bypassing lever  78 . A downward recess  97  is formed at a portion of center section  10  in the vicinity of motor mounting surface  10   m . A bottom end portion of camshaft  77  is inserted into recess  97 . In recess  97 , a bottom end portion of camshaft  77  is partly notched so as to form a cam  80 . 
     Center section  10  is bored with a diametrically small through-hole in parallel to motor shaft  22  between motor mounting surface  10   m  and recess  97 . A pin  79  is reciprocally disposed in the through-hole. One end of pin  79  is disposed close to cylinder block  24 . The other end of pin  79  is disposed close to cam  80  in recess  97 . 
     If the working vehicle provided with this transaxle apparatus  1  is towed by another working vehicle, lever  78  is rotated so as to rotate camshaft  77  integrally and make cam  80  push pin  79  so that the end of pin  79  projects from motor mounting surface  10   m  and pushes cylinder block  24 , thereby separating cylinder block  24  from motor mounting surface  10   m . Therefore, cylinder block  24  is hydraulically bypassed and each of hydraulic oil passages  111  and  112  is directly connected to the oil sump in housing  9 . Hydraulic motor  21  becomes rotatable freely from the hydraulic oil in hydraulic oil passages  111  and  112 . Consequently, axles  50 L and  50 R interlocking with motor shaft  22  become free so as to prevent resistance against the towing. 
     The hitherto descriptions are chiefly given on the structures of transaxle apparatus  1  which are common among some embodiments thereof described later. Now, description will be given on each distinctive control system of motor swash plate  23  of hydraulic motor  21 , namely, motor capacity control systems  200 ,  200 ′,  200 ′ a ,  200 ′ b  and  200 ′ c  which utilize hydraulic pressure of HST  8  for adjusting the speed reduction ratio of HST  8  correspondingly to load on engine  2 . 
     Previous to the descriptions thereof, the reason why the hydraulic pressure of at least either first hydraulic oil passage  11  or second hydraulic oil passage  112  is used as detection of the engine load is referred to. While the working vehicle travels, various resistances such as road resistance, air resistance, acceleration resistance, slope resistance and so on are generated on axles  50 L and  50 R. These resistances are transmitted as a torque against the driving force of motor shaft  22  through gear train  30  to motor shaft  22  of hydraulic motor  21 . 
     The greater the total resistance generated from axles  50 L and  50 R becomes, the greater hydraulic pressure force is required to drive motor shaft  22 . If the working vehicle drives forward, the hydraulic pressure of first hydraulic oil passage  111  is increased so much as to overcome the total resistance, thereby becoming higher than its essential pressure determined by setting the capacity of hydraulic pump  11  with pump swash plate  13 . 
     On the other hand, the increase of the total resistance results in the increase of load applied on engine  2 . 
     Consequently, when the hydraulic pressure of either first hydraulic oil passage ill or second hydraulic oil passage  112  is increased, the load on engine  2  is increased. 
     Thus, according to the detection of the increase of hydraulic pressure of the hydraulic oil passage of HST  8  regarded as the increase of load on engine  2 , the capacity of hydraulic motor  21  is automatically increased so as to increase the speed reduction ratio of HST  8 , thereby overcoming the load of engine  2 . 
     In transaxle apparatus  1  shown in FIGS. 1 to  6 , for example, a motor capacity control system  200  is provided for controlling the position of motor swash plate  23 , thereby controlling the capacity of hydraulic motor  21 . In this regard, motor capacity control system  200  consists of a hydraulic actuator moving motor swash plate  23 , load-detection means  201 , and actuator-control means  202 . In this embodiment, the hydraulic actuator is hydraulic cylinder  35 . Load-detection means  201  detects the hydraulic pressure of first hydraulic oil circuit  111  replacing the load on engine  2 . Actuator-control means  202  hydraulically controls hydraulic cylinder  35  on basis of the hydraulic pressure condition detected by load-detection means  201 . 
     Incidentally, there is no assumption that the working vehicle having transaxle apparatus  1  shown in FIGS. 1 to  6  travels backward with engine  2  under heavy load. Hence, transaxle apparatus  1  of this embodiment utilizes the detection of hydraulic pressure of only first hydraulic oil passage  111 , without second hydraulic oil passage  112 , replacing the detection of engine load. 
     However, hydraulic pressure of second hydraulic oil passage  112  may be used as detection of the engine load and for controlling motor swash plate  23 , additionally or alternatively to first hydraulic oil passage  111 , if it is required to overcome the engine load in backward traveling and if there is a room for such a device in transaxle apparatus  1 . 
     Now, referring to arrangement of motor swash plate  23  in transaxle apparatus  1  shown in FIGS. 1 to  6 , a round fulcrum shaft  99  serving as a fulcrum of movable motor swash plate  23  in its slanting is vertically disposed while being partly inserted into partition wall  9   i  of housing  9 . Motor swash plate  23  is vertically formed on its back surface with a sectionally half-round groove in correspondence to fulcrum shaft  99 . Motor swash plate  23  is arranged between upper and lower housing parts  9   a  and  9   b  so as to abut against heads of pistons  25 . Fulcrum shaft  99  is engaged in the groove of motor swash plate  23 . In this arrangement, motor swash plate  23  is rotated around fulcrum shaft  99  while the surface-of the groove slides against the peripheral surface of fulcrum shaft  99 . 
     Partition wall  9   i  is formed with a pair of contact surfaces  9   ia  and  9   ib  which are arranged oppositely to each other with respect to motor shaft  22  when viewed in plan as shown in FIGS. 5 and 6. Motor swash plate  23  is rotated between the minimum slant angle position with angle A 1  and the maximum slant position with angle A 2 . When motor swash plate  23  is located at the minimum slant angle position, motor swash plate  23  abuts against contact surface  9   ia , as shown in FIG.  5 . When motor swash late  23  is located at the minimum slant angle position, motor swash late  23  abuts against contact surface  9   ib , as shown in FIG.  6 . 
     As shown in FIGS. 4 to  6 , a control shaft  82  for rotating motor swash plate  3  is supported by housing  9  vertically in parallel to fulcrum shaft  99 , and in opposite to fulcrum shaft  99  with respect to motor shaft  22 . Control shaft  82  is formed at its intermediate portion into an eccentric half-round shaped cam  82   a . On the other hand, motor swash plate  23  is integrally formed on one lateral side thereof with an operation arm  23   a . Cam  82   a  of control shaft  82  is brought into contact with a vertical surface of operation arm  23   a  against the pressure force of pistons  25 . When control shaft  82  is rotated, cam  82   a  is revolved centering on the axis of control shaft  82  while abutting against operation arm  23   a , thereby moving motor swash plate  23 . 
     Control shaft  82  projects upwardly from housing  9  so as to be fixedly provided thereon with a control lever  83 . Lever  83  is pivotally connected with a piston rod of a hydraulic cylinder  35  serving as a hydraulic actuator for moving motor swash plate  23 . Hydraulic cylinder  35  is also pivoted on a vehicle frame. 
     As shown in FIG. 2, an oil port sleeve  34  is provided through a bottom wall of housing  9  and attached onto the bottom surface of center section  10  so as to be connected to first hydraulic oil passage  111  in center section  10 . Hydraulic oil passage  111  is higher-pressured when the working vehicle is driven forward. Hydraulic cylinder  35  is fluidly connected with first hydraulic oil passage  111  through an external pipe fitting (not shown) and oil port sleeve  34  as shown in FIG.  1 . 
     As shown in FIG. 1, hydraulic cylinder  35  is provided with a spring  84  biasing its piston rod in the contracting direction. When the hydraulic pressure of first hydraulic oil passage  111  is small (whether first hydraulic oil passage  111  is higher-pressured or depressed), the piston rod of hydraulic cylinder  35  is contracted by the force of spring  84  so that motor swash plate  23  abuts against first contact surface  9   ia , as shown in FIG. 5, wherein a surface of motor swash plate  23  abutting against pistons  25  is slanted at minimum angle A 1  from plane P. 
     As the hydraulic pressure of first hydraulic oil passage  111  is increased, a part of pressured oil in first hydraulic oil passage  111  flows into hydraulic cylinder  35  so as to extend the piston rod of hydraulic cylinder  35  and rotate lever  83  as an arrow x shown in FIG.  5 . 
     Accordingly, cam  82   a  of control shaft  82  retreats so that motor swash plate  23  with operation arm  23   a  abutting against cam  82   a  is naturally further slanted from angle A 1  by effect of the pressure of pistons  25 , thereby increasing the capacity of hydraulic motor  21 . Motor swash plate  23  is finally balanced when the hydraulic pressure of first oil passage  111  and the spring force of spring  84  become equal in hydraulic cylinder  35 . The balancing position of motor swash plate  23  is determined according to the hydraulic pressure of first oil passage  111 . 
     If the hydraulic pressure of first hydraulic oil passage  111  exceeds a certain degree, motor swash plate  23  comes to abut against contact surface  9   ib , as shown in FIG. 6, wherein the surface of motor swash plate  23  abutting against pistons  25  is slanted from plane P at maximum angle A 2  that is larger than angle A 1 . 
     Minimum angle A 1  and maximum angle A 2  of motor swash plate  23  of hydraulic motor  21  are desirably set to appropriate degrees in consideration of the permissible load range of engine  2 , the use of the working vehicle having this transaxle apparatus  1 , the capacity of hydraulic pump  11 , the speed reduction ratio of gear train  30  and so on. 
     In hydrostatic vehicle driving system shown in FIG. 1, which includes transaxle apparatus  1  shown in FIGS. 2 to  6 , hydraulic cylinder  35  is provided as the hydraulic actuator for moving motor swash plate  23 . Oil port sleeve  3   d  and the pipe fitting interposed between first hydraulic oil circuit  111  and hydraulic cylinder  35  serve as both load-detection means  201  and actuator-control means  202 . Thus, motor capacity control system  200  is structured. 
     Next, a hydrostatic vehicle driving system for a working vehicle according to the present invention shown in FIG. 7, which includes transaxle apparatus  1  provided with an alternative motor capacity control system  200 ′ as shown in FIGS. 8 to  12 , will be described. In FIGS. 7 to  12 , the same reference numerals designate identical or substantially similar parts or assemblies with those in FIGS. 1 to  6 . 
     Except matters peculiar to this embodiment described as follows, the hydrostatic vehicle driving system shown in FIG.  7  and transaxle apparatus  1  shown in FIGS. 8 to  12  have common structures with those shown in FIGS. 1 to  6 . 
     In this transaxle apparatus  1 , a piston  120  provided in housing  9  serves as a hydraulic actuator for rotating motor swash plate  23  of hydraulic motor  21 , constituting motor capacity control system  200 ′. 
     As shown in FIGS. 10 and 11, the back of motor swash plate  23  is formed into an arcuate convex. A retainer  109  is fittingly sandwiched between upper and lower housing parts  9   a  and  9   b  so as to slidably supporting motor swash plate  23 . Retainer  109  has a recessed arcuate surface  109   a . The arcuate surface of the back convex of motor swash plate  23  fittingly and slidably abuts against recessed arcuate surface  109   a  so that motor swash plate  23  is slanted while being guided by retainer  109 . 
     In this embodiment, the rotary axis of motor shaft  21  is placed on a joint surface between upper and lower housing parts  9   a  and  9   b . If retainer  109  for supporting motor swash plate  23  were to be molded by housing parts  9   a  and  9   b , it would be required that upper and lower housing parts  9   a  and  9   b  are formed with arcuate surfaces, respectively, and both the arcuate surfaces of upper and lower housing parts  9   a  and  9   b  are accurately joined with each other without slipping. Such accurate formation of arcuate surfaces on respective housing parts  9   a  and  9   b  is difficult. 
     From this view point, retainer  109  is made separately from housing parts  9   a  and  9   b . Even if there is unevenness between upper and lower housing parts  9   a  and  9   b , retainer  109  having recessed arcuate surface  109   a  is disposed across the joint surface between upper and lower housing parts  9   a  and  9   b  so as to slidably support motor swash plate  23  properly without abrasion. 
     This structure is typically applicable in the case that the rotary axis of hydraulic motor  21  is arranged in parallel to or coincidentally with the joint surface of a divisible housing. It may also be used for supporting a movable swash plate of a hydraulic pump whose rotary axis is disposed in parallel to or coincidentally with the joint surface of a divisible housing. 
     As shown in FIGS. 9 to  12 , center section  10  is additionally formed with a first extension portion  10   a  and a second extension portion  10   b . First extension portion  10   a  is extended downwardly from motor mounting surface  10   m . Second extension portion  10   b  is extended horizontally in parallel to motor shaft  22  from first extension portion  10   a  below hydraulic motor  21 . An utmost end of second extension portion  10   b  is located close to motor swash plate  23 . 
     A vertical oil hole  105  is formed in first extension portion  10   a . A horizontal cylinder chamber  106  is formed in second extension portion  10   b  so as to be connected with oil hole  105  and open toward motor swash plate  23 . Motor swash plate  23  is integrally provided with a plane operation arm  23   a  which is extended downwardly from the bottom end of motor swash plate  23 . Piston  120  consisting of a sleeve  121  and a ball joint  122  is interposed between cylinder chamber  106  and operation arm  23   a  of motor swash plate  23 . 
     Cylinder chamber  106  is provided at its intermediate portion with a step  106   a . Sleeve  121  is fittingly and slidably reciprocally inserted into a diametrically larger portion of cylinder chamber  106  between its open end and step  106   a . Ball joint  122  having a spherical portion and a plane portion is interposed between sleeve  121  and operation arm  23   a  while the spherical portion is fittingly and slidably inserted into sleeve  121  and the plane portion abuts against operation arm  23   a . When sleeve  121  abuts against step  106   a  as shown in FIG. 10, the slant angle of motor swash plate  23  from plane P perpendicular to motor shaft  22  is minimum angle A 1 . 
     Alternatively, cylinder chamber  106  may be provided in an extension portion formed by charge pump casing  16   a  instead of extension portions  10   a  and  10   b  of center section  10 . 
     In opposite to piston  120 , a return piston  88  is interposed between operation arm  23   a  and retainer  109 . In detail, retainer  109  is formed with a piston hole  125 . Return piston  88  consisting of a sleeve  85  and a ball joint  86  has a similar construction with piston  120 . Sleeve  85  is fittingly and slidably reciprocally inserted into piston hole  125  through a spring  84 . Ball joint  86  having a spherical portion and a plane portion is interposed between sleeve  85  and operation arm  23   a  while the spherical portion is fittingly and slidably inserted into sleeve  85  and the plane portion abuts against operation arm  23   a . Consequently, return piston  88  biases motor swash plate  23  with the force of spring  84  against piston  120 . 
     Piston hole  125  is further extended through housing  9  (lower housing part  9   b ) and open at an outer side wall of housing  9 . Within the wall of housing  9 , piston hole  125  is female-screwed and a threaded adjusting rod  87  is screwed into piston hole  125  and fastened with housing  9  through a nut  87   a . Adjusting rod  87  is rotated and axially moved by rotating nut  87   a  so as to adjust the spring force of spring  84 , thereby adjusting the initial position of motor swash plate  23 . 
     In retainer  109 , piston hole  125  is provided with a step  125 . When sleeve  85  abuts against step  125   a  as shown in FIG. 11, the slant angle of motor swash plate  23  from plane P is maximum angle A 2 . 
     Due to such a structure, when the hydraulic pressure of first hydraulic oil passage  111  is increased, the hydraulic pressure in cylinder chamber  106  in connection with first hydraulic oil passage  111  is increased so as to thrust piston  120  outward, thereby pushing operation arm  23   a  of motor swash plate  23 . The slant angle of motor swash plate  23  is increased so as to increase the capacity of hydraulic motor  21 . 
     On the other hand, return piston  88  is pushed by operation arm  23   a  so as to press spring  84 , thereby increasing the spring force of spring  84 . Consequently, motor swash plate  23  is balanced when the pressure force of piston  120  by hydraulic pressure of first oil passage  111  and the spring force of spring  84  become equal. The balancing position of motor swash plate  23  is determined according to the hydraulic pressure of first oil passage  111 . 
     In brief, piston  120  is operated according to the detected hydraulic pressure of first hydraulic oil passage  111  which is regarded as load on engine  2 . Accordingly, if the increase of hydraulic pressure of first hydraulic oil passage  111  is detected, motor swash plate  23  is rotated from minimum angle A 1  toward maximum angle A 2 . 
     The stationary or balancing position of motor swash plate  23 , in relation to the hydraulic pressure of first oil passage  111 , will be described in accordance with FIGS. 9 and 10. 
     When first oil passage  111  is hydraulically non-pressured or negatively pressured, piston  120  is detracted so that sleeve  121  abuts against step  106   a . At this time, motor swash plate  23  is located at the minimum slant angle position, wherein motor swash plate  23  is slanted at angle A 1  from plane P as shown in FIG.  10 . When motor swash plate  23  is set at the minimum slant angle position, hydraulic motor  21  has the minimum capacity so that the speed reduction ratio of HST  8  is the minimum if the position of pump swash plate  13  of hydraulic pump  11  is fixed. 
     The minimum slant angle position (or angle A 1 ) of motor swash plate  23  is determined according to the position of step  106   a . In this hydraulic pressure condition of first hydraulic oil passage  111 , it is possible to say that little or almost no load is applied on engine  2 . Therefore, the position of step  106   a  may be determined so as to set angle A 1  of motor swash plate  23  to the suitable minimum angle without consideration of the load on engine  2 , thereby determining the minimum capacity of hydraulic motor  21 . 
     As first hydraulic oil passage  111  is positively hydraulically pressured for driving the working vehicle forward, the hydraulic pressure of first hydraulic oil passage  111  makes piston  120  project out as an arrow y shown in FIG. 10 so as to push operation arm  23   a . Thus, the slant angle of motor swash plate  23  is increased and piston  88  is retracted. Finally, motor swash plate  23  is balanced between the thrusting force of piston  120  and the biasing force of spring  84  so as to become stationary. This balancing position of motor swash plate  23  is steplessly moved in proportion to the increase of hydraulic pressure of first hydraulic oil passage  111 . 
     When first hydraulic oil passage  111  is hydraulically pressured to a certain degree, sleeve  85  of piston  88  comes to abut against step  125   a  in piston hole  125 , thereby stopping motor swash plate  23  at the maximum slant angle position wherein motor swash plate  23  is slanted at angle A 2  from plane P, as shown in FIG.  11 . When motor swash plate  23  is set at the maximum slant position, hydraulic motor  21  has the maximum capacity so that the speed reduction ratio of HST  8  is the maximum if the position of pump swash plate  13  of hydraulic pump  11  is fixed. Even if the hydraulic pressure of first hydraulic oil passage  111  is increased beyond the certain degree, motor swash plate  23  is still stationary at the maximum slant angle position. 
     In this high-pressured condition of first hydraulic oil passage  111 , it is possible to say that heavy load is applied on engine  2 . Therefore, the position of step  125   a  should be determined so as to set angle A 2  of motor swash plate  23  to the suitable maximum angle under consideration of the maximum permissible load of engine  2 , thereby determining the maximum capacity of hydraulic motor  21 . 
     Incidentally, in association with arrangement of extension portions  10   a  and  10   b  below center section  10 , instead of inner suction filter  17  as shown in FIG. 3, a suction filter  17 ′ for filtering oil supplied to hydraulic oil passages  111  and  112  of HST  8  is externally attached onto a side wall of lower housing part  9   b , as shown in FIG.  12 . As shown in FIG. 9, suction filter  17 ′ is arranged in a dead space below control arm  60 , thereby minimizing transaxle apparatus  1 . Additionally, as shown in FIG. 11, a suction port  127  for suction filter  17 ′ is provided in a side wall of housing  9 , and a pipe  101  is interposed through a side wall of housing  9  between a discharge port of suction filter  17 ′ and a suction port  16   b  formed in charge pump casing  16   a  of charge pump  16 . 
     Therefore, oil in the oil sump of housing  9  is introduced into suction filter  17 ′ through suction port  127 , and sent to suction port  16   b  of charge pump  16  through a pipe  101 . As shown in FIG. 12, in charge pump casing  16   a  is formed a discharge oil passage  128  extending from discharge port  16   c . An oil groove  129  is formed between center section  10  and charge pump casing  16   a . Charge pump  16  discharges oil from discharge port  16   b  and discharge oil passage  128  to charge oil passage  93  in center section  10  through oil groove  129 . A relief valve  76  is connected to discharge oil passage  128  so as to regulate a charge pressure. 
     Motor capacity control system  200 ′ including piston  120  is drawn as a diagram in FIG.  7 . In this system  200 ′, the oil passage consisting of oil hole  105  and cylinder chamber  106  serve as load-detection means  201 , which detects the hydraulic oil pressure of first hydraulic oil passage  111 . Actuator-control means  202  for controlling Piston  102  is constituted by cylinder chamber  106 , return piston  88 , spring  84  and the like. 
     Next, various modifications of the hydrostatic vehicle driving system shown in FIG. 7, which includes motor capacity control system  200 ′ having piston  120  as a hydraulic actuator for moving motor swash plate  23  of hydraulic motor  21 , will be described in FIGS. 13 to  34 . 
     However, as long as possible, the distinction of each following modification of this hydrostatic vehicle driving system having motor capacity control system  200 ′ may be adapted to the first hydrostatic vehicle driving system including motor capacity control system  200  as shown in FIG. 1, which includes hydraulic cylinder  35  as a hydraulic actuator for motor swash plate  23 . 
     In FIG. 7, movable motor swash plate  23  is steplessly adjusted with its slant angle between angles A 1  and A 2 . Alternatively, movable motor swash plate  23  may be positionally switched between minimum angle A 1  and maximum angle A 2  as shown in FIGS. 13 and 14. 
     Referring to FIG. 13, a two-position switching valve  102  is interposed between first hydraulic oil passage  111  and piston  120 . Valve  102  is operated by inner pilot pressure. When hydraulic pressure of first hydraulic oil passage  111  exceeds the force of a return spring  103  of valve  102 , valve  102  is switched so as to thrust out piston  120  so that motor swash plate  23  is switched to the maximum slant position from the minimum slant position. 
     In such a structured motor capacity control system  200 ′ a , a pilot oil path of valve  102  and an oil path interposed between the hydraulic actuator (piston  120 ) and first hydraulic oil passage  111  serve as load detection means  201 . Valve  102 , spring  103  and the like serve as actuator-control means  202  which controls the position of piston  120  in accordance with the hydraulic pressure replacing load on engine  2 , detected by load-detection means  201 . 
     Referring to FIG. 14, piston  120  is fluidly connected with discharge port  16   c  of charge pump  16  through two-position switching valve  102 . Pilot pressure of valve  102  is led from first hydraulic oil passage  111 . 
     If the pilot pressure force from first hydraulic oil passage  111  exceeds the force of spring  103  of valve  102 , valve  102  is switched so that a part of oil discharged from charge pump  16  is supplied to the hydraulic actuator (piston  120 ) so as to thrust out piston  120 , whereby motor swash plate  23  is switched to the maximum slant position from the minimum slant position. 
     Additionally, valve  102  is provided with a manual operation lever  102   a . When the working vehicle is at work, an operator may operate lever  102   a  so as to fix valve  102  at the position for holding motor swash plate  23  at the maximum slant position while engine  2  may be subjected to heavy load, thereby keeping the maximum capacity of hydraulic motor  21  and enabling the working vehicle to drive at a constant low speed. 
     In this motor capacity control system  200 ′ b , a pilot oil path of valve  102  serves as load-detection means  201 . Valve  102 , spring  103  and the like serve as actuator-control means  202  which controls piston  120  in accordance with the detected hydraulic pressure replacing load of engine  2 . 
     A relief valve  104  determines the hydraulic pressure of piston  120 . The excessive pressured oil for the hydraulic actuator is released to charge oil passage  93  through relief valve  104  so as to compensate inner oil leak of HST  8 . 
     The working vehicle having transaxle apparatus  1  may require that the capacity of variable displacement hydraulic motor  21  is fixed to the maximum in some cases, as being referred to in FIG.  14 . For example, if the working vehicle is a lawn tractor, the capacity of hydraulic motor  21 , which is desirably switched between the smaller and the greater during regular traveling of the working vehicle, may be desired to be fixed to the greater for overcoming engine load during its lawn-mowing. 
     Considering such a case, referring to FIGS. 15 to  19 , center section  10  is optimally formed therein with a piston drive oil passage  140 , which is extended from the discharge port of charge pump  16  through relief valve  104  to charge oil passage  93  among check valves  19  and the pair of check valves  26 . 
     The oil discharged from charge pump  16  and led into oil Passage  140  is pressured sufficiently for holding motor swash plate  23  at maximum angle A 2 , if it is supplied in cylinder chamber  106 . 
     A first two-position switching valve  131  and a second two-position switching valve  132  are fluidly connected with piston drive oil passage  140  in parallel. 
     In this motor capacity control system  200 ′ c , piston drive oil passage  140  and two valves  131  and  132  serve as actuator-control means  202 . For constituting load-detection means  201 , a shuttle valve  141  is interposed between first and second hydraulic oil circuits  111  and  112 , and a pilot oil passage  145  is extended from shuttle valve  141 . 
     First valve  131  is a hydraulic pilot valve, which is operated by pilot oil through pilot oil passage  145  from shuttle valve  141 . First valve  131  has three ports. A port A is connected to oil passage  140 . A port B is a drain port. A port C is connected to second valve  132  through a connection oil passage  135 . 
     When the hydraulic pressure in both hydraulic oil passage  111  and  112  is less than a pilot pressure for first valve  131 , in first valve  131 , ports B and C are connected with each other so as to drain oil in connection oil passage  135  and port A is shut from both ports B and C in first valve  131 , as shown in FIG.  16 . 
     When the pilot pressure for first valve  131  arises in either hydraulic oil passage  111  or  112 , in first valve  131 , ports A and C are connected with each other so as to connect oil passage  140  with connection oil passage  135  and port B is shut from both ports A and C in first valve  131 , as shown in FIG.  17 . 
     Second valve  132  is mechanically switched between two positions by link with a mode selection lever  36 , which is provided beside a driver&#39;s seat and manipulated between “a work mode position m 1  ” and “a regular traveling mode position m 2 ”. Second valve  132  may alternatively be a solenoid valve which is switched on/off by detecting the position of lever  36 . Second valve  132  also has three ports. A port A is connected to oil passage  140 . A port B is connected to first valve  131  through connection oil passage  135 . A port C is connected to cylinder chamber  106  through oil passage  105 ′. 
     When mode selection lever  36  is located at work mode position m 1 , in second valve  132 , ports A and C are connected with each other so as to connect oil passage  140  to cylinder chamber  106  and port B is shut from both ports A and C, as shown in FIGS. 16 and 17. Therefore, regardless the position of first valve  131 , certain pressured oil discharged from charge pump  16  is supplied into cylinder chamber  106  through oil passage  140 , second valve  132  and oil passage  105 ′, so that piston  120  is thrust out against spring  84  so as to rotate motor swash plate  23  to maximum angle A 2 , thereby fixing the capacity of hydraulic motor  21  to the maximum. 
     When mode selection valve  36  is located at regular traveling position m 2 , in second valve  132 , ports B and C are connected with each other so as to connect oil passage  135  to cylinder chamber  106  and port A is shut from both ports B and C, as shown in FIGS. 18 and 19. 
     In this condition, if first valve  131  is free from pilot pressure, as shown in FIG. 18, oil is drained from cylinder chamber  106  through second and first valves  132  and  131  while piston  120  is retracted by biasing force of spring  84  so as to locate motor swash plate  23  at minimum angle A 1 , thereby establishing the minimum capacity of hydraulic motor  21 . 
     On the other hand, while mode selection lever  36  being located at regular traveling position m 2 , if first valve  131  is operated by pilot pressure, as shown in FIG. 19, oil is supplied from oil passage  140  into cylinder chamber  106  through first and second valves  131  and  132  so as to thrust out piston  120  against the biasing force of spring  84 , thereby locating motor swash plate  23  at maximum angle A 2  so as to establish the maximum capacity of hydraulic motor  21 . 
     Incidentally, in this embodiment, the capacity of hydraulic motor  21  can be set to the maximum, whether the working vehicle may travel forward or backward, because motor swash plate  23  is rotated to maximum angle A 2  in correspondence to the increase of hydraulic pressure in either first or second hydraulic oil circuit  111  or  112 . 
     If the working vehicle having such a structure shown in FIGS. 15 to  19  is a lawn tractor, the work mode means its lawn-mowing. During the lawnmowing, the working vehicle must drive at a constant low speed for preventing unevenness of the mowed lawn, however great resistance may be generated against driving axles  50 L and  50 R. 
     Then, in above-described accelerator system  100 , which has no relation to speed control pedal  27 , accelerator lever  20  with friction lock means  45  is set to fix the output rotary speed of engine  2  optimally. Mode selection lever  36  is located at work mode position m 1 , as shown in FIGS. 16 and 17, so as to fix motor swash plate  23  at maximum angle A 2 , thereby establishing the maximum capacity of hydraulic motor  21  so as to overcome the total resistance applied on axles  50 L and  50 R and avoid overload on engine  2 . 
     In such a precondition, speed control pedal  27  is pressed so as to establish optimal capacity of hydraulic pump  11 . While speed control pedal  27  is fixed at its pressed position, the greatest speed reduction ratio of HST  8  is fixed, whereby the working vehicle can cruise at constant low speed. 
     If the working vehicle is workless and going to travel on road, it may be possible that the working vehicle travels at a desirable high speed, repeats frequent stopping and starting, or climbs a slope. Thus, it is desired that the capacity of hydraulic motor  21  is switched between the smaller and the greater in correspondence to the fluctuations of the total resistance against driving axles  50 L and  50 R. 
     In other words, it is preferred that the capacity of hydraulic motor  21 , reduced for smooth traveling under little resistance, is increased only when axles  50 L and  50 R are subjected to great resistance. 
     Then, accelerator lever  20  with friction lock means  45  is set to fix the output rotary speed of engine  2  optimally. Mode selection lever  36  is located at regular traveling mode position m 2 , as shown in FIGS. 18 and 19. 
     In such a precondition, speed control pedal  27  is pressed to an optimal depth so as to establish an optimal traveling speed. While the resistance against driving axles  50 L and  50 R is small and the hydraulic pressure in both first and second hydraulic oil circuits  111  and  112  is less than the pilot pressure for first valve  131 , oil passage  140  is shut from cylinder chamber  106  by first valve  131  as shown in FIG. 18, thereby detracting piston  120  so as to keep motor swash plate  23  at minimum angle A 1 , whereby the highest possible traveling speed can be attained while speed control pedal  27  is pressed to a certain depth. Thus, working vehicle can accelerate and cruise desirably. 
     If the resistance against driving axles  50 L and  50 R is increased so that a pilot pressure for first valve  131  rises in either first or second hydraulic oil circuit  111  or  112 , oil passage  140  is brought into communication with cylinder chamber  106  by first valve  131  as shown in FIG. 19, thereby thrusting out piston  120  so as to locate motor swash plate  23  at maximum angle A 2 , whereby the lowest possible traveling speed is attained while speed control pedal  27  is pressed to a certain depth. Thus, the resistance can be overcome so as to avoid overload on engine  2 . 
     Referring to an embodiment shown in FIGS. 20 to  23 , there is adopted motor capacity control system  200 ′ c  identical with that shown in FIGS. 16 to  19 . The distinctive point of this embodiment from that shown in FIGS. 16 to  19  is that link mechanism  300  is interposed between accelerator system  100  and speed control pedal  27 . Link mechanism  300  brings throttle arm  134  of carburetor  130  into connection with speed control pedal  27  according to the manipulation of mode selection lever  36 . This structure and effect is described. 
     A first clutch shaft  53  is fixed to control arm  61  so as to be rotated integrally with control arm  61 . A clutch slider is axially slidably provided around first clutch shaft  53  through a spline. A second clutch shaft  55  is disposed coaxially with first clutch shaft  53 . 
     A clutch  54  including the clutch slider on first clutch shaft  53  is interposed between first and second clutch shafts  53  and  55 . 
     The clutch slider interlocks with mode selection lever  36  so as to be slid along first clutch shaft  53  by rotating mode selection lever  36 . Accordingly, clutch  54  is disengaged when mode selection lever  36  being located at work mode position m 1  as shown in FIGS. 20 and 21, and is engaged when mode selection lever  36  being located at regular traveling mode position m 2  as shown in FIGS. 22 and 23. 
     The other end of second clutch shaft  55  is fixed to a center position of an arm  57 . Arm  57  has two opposite extended portions from its center point in connection with second clutch shaft  55 . A cam plate  58  is disposed adjacent to arm  57 . A cam plate  58  is pivoted at its one end and formed with a pair of concentrically arcuate first and second slots  58   b  and  58   c  centering on a pivot point  58   a  of cam plate  58 . 
     A pair of first and second links  62  and  63  are interposed between arm  57  and cam plate  58 . One end of each of links  62  and  63  is connected to each of the opposite extended portions of arm  57 . The other end of each of links  62  and  63  is slidably inserted into each of first and second slots  58   b  and  58   c , respectively. First slot  58   b  and the end of first link  62  therein are nearer to pivot point  58   a  than second slot  58   c  and the end of second link  63  therein. A clearance of first slot  58   b  for play of first link  62  does not need to be so long as that of second slot  58   c  for the same amount of play of second link  63 . Thus, first slot  58   b  may be made shorter than second slot  58   c . This effect is discussed later. 
     When mode selection lever  36  is located at work mode position m 1  so as to disengage clutch  54  as shown in FIGS. 20 and 21, arm  57  stays at its neutral position so that the end of each of links  62  and  63  is held at one end of each of slots  58   b  and  58   c . Even if mode selection lever  36  is located at regular traveling mode position m 2  so as to engage clutch  54  as shown in FIGS. 22 and 23, when speed control pedal  27  is unpressed, arm  57  still stays at the neutral position so that the end of each of links  62  and  63  is still held at the one end of each of slots  58   b  and  58   c.    
     A wire  59  is interposed between cam plate  58  and throttle arm  134 . As mentioned above, throttle arm  134  is connected to accelerator lever  20  with friction lock means  45  through wire  64 . In this embodiment, friction lock means  45  is switched between a locking state and an unlocking state. Mode selection lever  36  also interlocks with friction lock means  45 . As shown in FIGS. 20 and 21, when mode selection lever  36  is located at work mode position m 1 , accelerator lever  20  is locked with friction lock means  45  so as to fix the output rotary speed of engine  2 . As shown in FIGS. 22 and 23, when mode selection lever  36  is located at regular traveling mode position m 2 , accelerator lever  20  is unlocked from friction lock means  45 , whereby accelerator lever  20  returns to its neutral position after its being released from manipulation force. 
     Due to such a construction as shown in FIGS. 20 to  23 , when mode selection lever  36  is located at work mode position m 1 , throttle arm  134  is free from speed control pedal  27  because clutch  54  is disengaged. 
     Therefore, the throttle of carburetor  130  is controlled only by accelerator lever  20  regardless of the pressing of speed control pedal  27 . Also, manipulated acceleration lever  20  is locked with friction lock means  45 , thereby enabling the output rotary speed of engine  2  to be fixed. In this condition, the maximum capacity of hydraulic motor  21  is kept whether pilot pressure for first valve  131  may rise or not in one of first and second hydraulic oil circuits  111  and  112 , as shown in FIGS. 20 and 21. Consequently, the working vehicle securely cruises at a constant speed established by pressed speed control pedal  27  however great resistance is generated against axles  50 L and  50 R. 
     When mode selection lever  36  is located at regular traveling mode position m 2  as shown in FIGS. 22 and 23, clutch  54  is engaged so as to unify both first and second clutch shafts  53  and  55 , thereby interlocking speed control pedal  27  with throttle arm  134 . 
     In this state, when fore pedal portion  27   a  of speed control pedal  27  is pressed, control arm  61  is rotated through link rod  51  so that arm  57  is rotated from its neutral position in a direction as an arrow z shown in FIGS. 22 and 23 together with control shaft  60 , first clutch shaft  53  and second clutch shaft  55 . Accordingly, cam plate  58  is pulled and rotated by arm  57  through first link  62  while second link  63  plays with its end sliding in second slot  58   c  of cam plate  58 . Rotated cam plate  58  pulls throttle arm  134  so as to widen the throttle of carburetor  130 , thereby accelerating the output rotation of engine  2 . 
     On the other hand, when rear pedal portion  27   b  of speed control pedal  27  is pressed, arm  57  is rotated from the neutral position oppositely to arrow z shown in FIGS. 22 and 23 together with control shaft  60 , first clutch shaft  53  and second clutch shaft  55 . Accordingly, cam plate  58  is pulled and rotated by arm  57  through second link  63  while first link  62  plays with its end sliding in first slot  58   b  of cam plate  58 . Rotated cam plate  58  pulls throttle arm  134  so as to widen the throttle of carburetor  130 , thereby accelerating the output rotation of engine  2 . 
     Also, accelerator lever  20  is unlocked from friction lock means  45 . Therefore, accelerator lever  20 , while being untouched, stays at the neutral (idling) position. Accordingly, during the regular traveling mode, the throttle of carburetor  130  is controlled only by speed control pedal  27 , in other words, speed control pedal  27  is pressed so as to change both the throttle of carburetor  130  and the capacity of hydraulic pump  11 . 
     The pulling stroke of first link  62  according to the rotation of arm  57  when control arm  61  being rotated to a certain degree from its neutral position by pressing fore pedal portion  27   a  is the same with that of second link  63  when control lever  61  being oppositely rotated to the same degree from the neutral position by pressing rear pedal portion  27   b . However, the clearance of first slot  58   b  for play of first link  62  while second link  63  being pulling cam plate  58  at a certain stroke is shorter than that of second slot  58   c  for play of second link  63  while first link  62  being pulling cam plate  58  at the same stroke. Thus, even if the capacity of hydraulic pump  11  is set to the same degree, the rotational degree of throttle arm  134  during backward traveling is less than that during forward traveling so that the acceleration of engine  2  is restricted during backward traveling in comparison with that during forward traveling. 
     Generally, while the working vehicle traveling backward, it is rare that the working vehicle accelerates suddenly or climbs a slope. Therefore, such a restriction of acceleration during backward traveling is reasonable and advantageous in fuel-saving. 
     Anyway, as shown in FIGS. 22 and 23, while mode selection lever  36  is located at regular traveling mode position m 2 , the capacity of hydraulic motor  21  is switched between the minimum and maximum according to detection of hydraulic pressure in either first or second hydraulic oil circuit  111  or  112  in the same manner as shown in FIGS. 18 and 19. 
     For application to the two embodiments of hydrostatic vehicle driving system shown in FIGS. 16 to  19  and FIGS. 20 to  23 , transaxle apparatus  1  may be modified as shown in FIGS. 24 to  26 . 
     In center section  10 , oil passage  105 ′ replacing vertical oil hole  105  is separated from both first and second hydraulic oil circuits  111  and  112  and downwardly open through a port sleeve  34   a  at the bottom of housing  9  so as to be connected to port C of second valve  132 . 
     Also, charge oil passage  93  vertically bored in center section  10  is downwardly open through a port sleeve  34   b  at the bottom of housing  9  so as to be connected to ports A of first and second valves  131  and  132 . 
     This structure of transaxle apparatus  1  is provided with such outward open ports for their connection with first and second valves  131  and  132  assumed to be provided out of housing  9 . However, assuming that first and second valves  131  and  132  are disposed in housing  9 , the downward open oil ports may be replaced with those open in housing  9 . 
     Description will be given on another hydrostatic vehicle driving system shown in FIGS. 27 to  34 . 
     The capacity of hydraulic motor  21  is controlled by the foregoing motor capacity control system  200 ′ c  including valves  131  and  132 . The capacity of hydraulic motor  21  is fixed to the maximum when mode selection lever  36  is located at work mode position m 1 . It is switched between the minimum and maximum when mode selection lever  36  is located at regular traveling mode position m 2 . 
     Also, in this hydrostatic vehicle driving system, the capacity of hydraulic pump It is controlled by pressing speed control pedal  27  during the work mode, and it is automatically controlled according to the control of output rotary speed (acceleration) of engine  2  during the regular traveling mode. 
     For the automatic control of the capacity of hydraulic pump  11  during the regular traveling mode, transaxle apparatus  1  of this embodiment is provided with an automatic speed control system  160  for moving pump swash plate  13  of hydraulic pump  11  in addition to the linkage from speed control pedal  27 . In this regard, as shown in FIG. 28, an orifice  170  is provided on the way of piston drive oil passage  140  extended form the discharge port of charge pump  16 . A first cylinder drive oil passage  171  is extended from the upstream of orifice  170  in oil passage  140 . A second cylinder drive oil passage  172  is extended from the downstream of orifice  170  in oil passage  140 . 
     While engine  2  drives, first cylinder drive oil passage  171  is hydraulically pressured higher than second cylinder drive oil passage  172  because of orifice  170 . Hydraulic pressure in piston drive oil passage  140  is increased in proportion of acceleration of engine  2  because oil passage  140  is supplied with oil discharged from charge pump  16  driven by pump shaft  12  which is driven synchronously with engine  2 . Consequently, the difference of hydraulic pressure between first and second cylinder drive oil passages  171  and  172  is increased in proportion to acceleration of engine  2 . 
     A pair of double-acting hydraulic cylinders, which are a forward drive cylinder  173  and a backward drive cylinder  174 , are provided. Each of oil passages  171  and  172  branches into two ways so as to be connected to respective cylinder chambers of each of cylinders  173  and  174  through a forward/backward travel switching valve  175 , which is a manual two-position switching valve interlocking with a manual forward/backward travel selection lever  28 ′. 
     Forward/backward travel switching valve  175  is a double valve consisting of a forward drive valve  175   a  and a backward drive valve  175   b , which are switched together by switching lever  28 ′. When forward/backward travel selection lever  28 ′ is located at a forward traveling position F, forward drive valve  175   a  supplies oil from both passages  171  and  172  into the respective chambers of forward drive cylinder  173 , and backward drive valve  175   b  drains oil from both chambers of backward drive cylinder  174 . When forward/backward travel selection lever  28 ′ is located at a backward traveling position R, backward drive valve  175   b  supplies oil from both passages  171  and  172  into the respective chambers of backward drive cylinder  174 , and forward drive valve  175   a  drains oil from both chambers of forward drive cylinder  173 . 
     Whether forward/backward travel selection lever  28 ′ may be located at forward traveling position F or backward traveling position R, the piston stroke of each of cylinders  173  and  174 , which is supplied with oil from forward/backward travel switching valve  175 , is increased in proportion to the increase of hydraulic pressure difference between oil passages  171  and  172 , that is, acceleration of engine  2 . 
     An arm  180  having three arm portions  180   a ,  180   b  and  180   c  is provided with its center portion pivoted. First and second arm portions  180   a  and  180   b  are oppositely extended from the center pivotal portion. Third arm portion  180   c  is extended from the center pivotal portion in perpendicular to first and second arm portions  180   a  and  180   b . A piston rod of forward driving cylinder  173  is connected to first arm portion  180   a , and that of backward driving cylinder  174  to second arm portion  180   b.    
     When forward/backward travel selection lever  28 ′ is located at forward traveling position F, backward drive cylinder  174  is free from hydraulic pressure and the piston rod of forward drive cylinder  173  pulls first arm portion  180   a  at a stroke as much as the difference of hydraulic pressure between oil passages  171  and  172 . When forward/backward travel selection lever  28 ′ is located at backward traveling position R, forward drive cylinder  173  is free from hydraulic pressure and the piston rod of backward drive cylinder  174  pulls second arm portion  180   b  at a stroke as much as the difference of hydraulic pressure between oil passages  171  and  172 . 
     Selection means  150  is interposed among control arm  61 , speed control pedal  27  and arm  180 . Through selection means  150 , control arm  61  is selectively connected to either swing arm  27   c  of speed control pedal  27  or third arm portion  180   c  of arm  180 . This switching of connection depends upon the location of mode selection lever  36 . When mode selection lever  36  is located at work mode position m 1 , speed control pedal  27  interlocks with control arm  61 . When mode selection lever  36  is located at regular traveling mode position m 2 , arm  180  interlocks with control arm  61 . In this state, the rotation of control arm  61  is synchronized with that of arm  180 . Accordingly, The rotational angle of control arm  61  corresponds to the stroke of each of cylinders  173  and  174 . The rotational direction of control arm  61  from its neutral position depends which of cylinders  173  and  174  is hydraulically pressured, that is, at which of positions F and R forward/backward travel selection lever  28 ′ is located. In this regard, all of FIGS. 28 to  32  show that forward/backward travel selection lever  28 ′ is located at forward traveling position F. When forward/backward travel selection lever  36  is located at backward traveling position R, forward drive valve  175   a  and backward drive valve  175   b  of forward/backward travel switching valve  175  are set at the other positions than those shown in FIGS. 28 to  32 . 
     Consequently, while mode selection lever  36  is located at work mode position m 1  as shown in FIGS. 28 and 31, control arm  61  interlocking with pump swash plate  13  of hydraulic pump  11  is rotated by pressing speed control pedal  27 , that is, the capacity of hydraulic pump  11  is controlled by pressing speed control pedal  27 . The discharge direction of hydraulic pump  11  depends upon which is pressed, fore pedal portion  27   a  or rear pedal portion  27   b . FIG. 31 shows that speed control pedal  27  is neutral. FIG. 31 shows that fore pedal portion  27   a  is pressed. 
     In this state, motor swash plate  23  is fixed at maximum angle A 2  so as to keep the maximum capacity of hydraulic motor  21  in the above mentioned manner using valves  131  and  132 , thereby overcoming load on engine  2  while being traveling at an optimal speed determined by pressing speed control pedal  27 . 
     On the other hand, as shown in FIGS. 29,  30  and  32 , while mode selection lever  36  is located at regular traveling mode position m 2 , control arm  61  is rotated by rotation of arm  180  in correspondence to the acceleration degree of engine  2 . 
     in this regard, as shown in FIGS. 27 to  32 , the working vehicle of this embodiment is provided with accelerator system  100 ′ as described above. This may be replaced with another such as accelerator system  100 . In this system  100 ′, accelerator lever  20  and momentary accelerator pedal  46  are used for controlling throttle arm  134  of carburetor  130 . Moreover, while mode selection lever  36  is located at regular traveling mode position m 2 , accelerator lever  20  and momentary accelerator pedal  46  are also used for controlling the capacity of hydraulic pump  11 . In this state, speed control pedal  27  is useless. FIGS. 29 and 32 show that engine  2  is neutral so that arm  180  stays at its initial position, thereby keeping HST  8  neutral. FIG. 31 shows that accelerator lever  20  or momentary accelerator pedal  46  is operated to some degree for acceleration of engine  2  so that arm  180  is rotated so much, thereby automatically determining the capacity of hydraulic pump  11  without pressing speed control pedal  27 . 
     Also, while mode selection lever  36  is located at regular travel mode position m 2 , motor swash plate  23  is switchable between minimum angle A 1  and maximum angle A 2  so as to vary the capacity of hydraulic motor  21  in correspondence to the load on engine  2 . FIGS. 29 and 31 shows that swash plate  23  is set at minimum angle A 1  while light load being applied on engine  2 . FIG. 32 shows that motor swash plate  23  is set at maximum angle A 2  under heavy load applied on engine  2 . 
     In this embodiment, mode selection lever  36  and its base portion serving as selection means  150  are structured such as shown in FIG. 34. A horizontal base shaft  123  is rotatably supported by an optimal portion of the working vehicle. Mode selection lever  36  is disposed substantially upwardly from base shaft  123 . A U-like shaped segment  36   a  is fixedly provided onto the bottom end of mode selection lever  36  so as to straddle base shaft  123 . A pin  124  penetrates base shaft  123  together with segment  36   a  so as to cross the axis of base shaft  123  perpendicularly, so that mode selection lever  36  is not rotatable around base shaft  123 , but is rotatable around pin  124  in the axial direction of base shaft  123 . 
     A first input arm  151  and a second input arm  152  are juxtaposed along base shaft  123  so as to sandwich mode selection lever  36 . First and second input arms  151  and  152  are formed with respective boss portions  151   a  and  152   a  which are rotatably disposed around base shaft  123 . 
     First input arm  151  is connected to swing arm  27   c  of speed control pedal  27  through an optimal linkage. Second input arm  152  is connected to third arm portion  180   c  of arm  180 . 
     Above boss portions  151   a  and  152   a , first and second input arms  151  and  152  are formed with respective guide plate portions  151   b  and  152   b . Guide plate portions  151   b  and  152   b  are notched toward mode selection lever  36  so as to form respective guide notches  151   c  and  152   c  facing each other. Guide notch  151   c  serves as work mode position m 1  for mode selection lever  36 . Guide notch  152   c  serves as regular traveling mode position m 2  for mode selection lever  36 . 
     An output arm  153  is tied through its boss portion  153   a  together with base shaft  123 . Output arm  153  fixed to base shaft  123  in such a manner is connected to control arm  61  interlocking with pump swash plate  13  of hydraulic pump  11 . 
     Therefore, when mode selection lever  36  is located at work mode position m 1 , that is, when mode selection lever  36  is engaged in guide notch  151   c , first guide arm  151  is integrated with base shaft  123  through mode selection lever  36  so as to connect speed control pedal  27  with control arm  61 . During this state, first input arm  151  is rotated together with base shaft  123  and output arm  153  around the axis of base shaft  123  in correspondence to the pressing of either pedal portion  27   a  or  27   b  of speed control pedal  27 , thereby rotating control arm  61  in connection with output arm  153  so as to rotate pump swash plate  13 . 
     On the other hand, when mode selection lever  36  is located at regular traveling mode position m 2 , that is, when mode selection lever  36  is engaged in guide notch  152   c , second guide arm  152  is integrated with base shaft  123  through mode selection lever  36  so as to connect arm  180  with control arm  61 . During this state, second input arm  152  is rotated together with base shaft  123  and output arm  153  around the axis of base shaft  123  in correspondence to the stroke of either hydraulic cylinder  173  or  174 , that is, the rotational angle of throttle arm  134  operated by accelerator lever  20  and/or momentary accelerator pedal  46  and the location of forward/backward travel selection lever  28 ′, thereby rotating control arm  61  so as to rotate pump swash plate  13 . 
     Transaxle apparatus  1  may be modified for this embodiment as shown in FIG.  33 . Similarly with transaxle apparatus  1  shown in FIGS. 24 to  26 , port sleeves  34   a  is provided for connecting cylinder chamber  106  in center section  10  with second valve  132  disposed out of housing  9 . Also, port sleeve  34   b  is provided for bringing charge oil passage  93  among check valve  19  and two check valves  26  into connection with first valve  131  and second valve  132  disposed out of housing  9 . 
     Furthermore, in this embodiment, center section  10  is provided with a part of piston drive oil passage  140  including orifice  170  (not shown). 
     A pair of port sleeves  34   c  and  34   d  are provided through the bottoms of center section  10  and housing  9  so as to be open downward, from which respective oil passages  171  and  172  are extended. 
     Although the invention has been described in its preferred form with a certain degree of particularity, it is understood that the present disclosure of the preferred form has been changed in the details of construction and the combination and arrangement of parts may be resorted to without departing from the spirit and the scope of the invention as hereinafter claimed.