A system for controlling a pump apparatus

Abstract: A torque control system for variable delivery pumps (2a)(2b) receiving torque from an engine employs a signal related to an engine torque available at each engine speed setting to control the torque, and thus the amount of engine power that is absorbed by the variable delivery pumps (2a)(2b). In a low-power regime, wherein the fuel injected into the engine is less than maximum, torque control is effected by comparing the available torque with an amount of torque that would be available at maximum fuel injection. In a high-power regime, wherein the fuel injected is at least a predetermined value, torque control is effected by comparing the available torque with an amount of torque that is available at a combination of a set engine speed (10) and an actual engine speed (11). A bias circuit biases the stable speed operating point a small amount below the maximum speed, whereby improve fuel economy is attained. In a system with more than one variable delivery pump, when less than the full number of pumps (2a)(2b) is employed, control is effected in the low-power regime, even though all pumps in service are receiving maximum fuel.

Description: The present invention relates to a system for controlling the torque of a pump of a type frequently used in construction machines and other similar machines in which regulators are driven by one or more pumps. Figure 6 is a schematic drawing showing a configuration of a conventional apparatus using two pumps. An engine 1 drives the variable delivery pumps 2a and 2b which include swash plates 3a and 3b, to vary the discharge flow rates. Swash plates 3a and 3b are driven by regulators 4a and 4b. An operating lever 5 provides a pilot pressure, that varies in proportion to its opening position, through lines 6a and 6b to the regulators 4a and 4b so that one or both of the swash plates 3a and 3b may be driven in dependence on the setting of operation lever 5. Swash plates 3a and 3b are also controlled according to the discharge pressure of the pumps 2a and 2b. A change of position of the swash plates 2 is limited so that the change of position does not cause the pump power demand to exceed the output power of the engine when load is applied, i.e. when discharge pressure is great. Referring now to Figure 3 which shows a characteristic engine output torque. Conventionally, the total output power of a pump is set to within a margin (in excess) of the engine design power output. This permits the engine to operate within a governor area. The governor area is the range of variation of the engine output torque capable of controlling the engine speed over a relatively narrow range. Point B is located at the upper limit of the governor area. Below Point B, the engine speed can be controlled by varying rack displacement, (i.e, controlling the fuel supply) as long as rack displacement remains less than the maximum rack displacement. At the maximum rack displacement, the speed is in a lagging area. Fuel consumption characteristic (in the figure, higher engine speed indicates less economical fuel consumption) is not favourable with the engine speed at Point B, suggesting that it is not operating efficiently. The output of each pump is a fixed value. Therefore, when only one of the two pumps is driven, less than half the engine power is used. The prior art presents the following drawbacks: (1) the engine is operated under unfavourable inefficient conditions with respect to fuel consumption, and (2) available engine power is not used to its full extent. Another example of a controlling device is disclosed in Japanese Patent Application Laid-Open No. 50686/1988 that calls for establishing a pump absorption (power demand) characteristic which makes a pump do a specified amount of work based on engine speed. This is accomplished by controlling the pump swash plates (in other words the pump discharge flow rate) according to the pump power demand or absorption characteristic and the pump discharge pressure. As shown in Figure 3, engine output characteristics vary considerably, depending on whether the engine speed is in the governor area, where the engine speed can be maintained at a more or less a fixed level regardless of a small change in engine output torque, or in the lagging area, where such control is ineffective. Therefore, while it is necessary to control the pump absorption torque in accordance with the area in which the rack displacement is located, the control device presented in the above Japanese Patent Application Laid-Open No. 50686/1988 does not solve this problem. Further guidance to the state of the art can be found by reference to GB-2 171 757, with reference to which claim 1 is characterised. It is an object of the present invention to provide a system for controlling a pump apparatus which alleviates the aforementioned technical problems of the prior art. It is a further object of the invention to provide a system which operates the engine under optimum conditions. Accordingly the present invention provides an apparatus for controlling the torque in an engine and pump system wherein the engine is coupled to drive a plurality of variable delivery pumps the engine having a fuel injection pump, comprising: an accelerating lever, a set engine speed device responsive to the accelerating lever, an actual engine speed sensor, an engine output torque calculator responsive to the set engine speed, characterised in that there is provided; a first correction circuit including; an under speed set means, a first summer, a first proportional integral controller to calculate a first correction torque according to the difference between a target engine speed and the actual engine speed, a second summer to correct the engine output torque by the first correction torque in the lagging area, a second correction circuit including; a rack command value generator responsive to the accelerator lever to generate a rack command value to control the fuel injection pump, a third summer, a second proportional-intregral controller to calculate a second correction torque according to the difference between an actual rack command value and a maximum rack command value, a summer to correct the engine output torque by the second correction torque in the governor area, a comparator which compares the rack command value with a predetermined rack command value related to the maximum rack command value to actuate a switch whereby the pump is controlled by the first correction circuit when the rack command value is greater than or equal to the predetermined rack command value, and by the second correction circuit when the rack command value is less than the predetermined value. In a low-power range, wherein the fuel injected into the engine is less than maximum, pump torque control is effected by comparing the available torque with a predetermined value of torque that would be available near maximum fuel injection. In a high-power regime, wherein the fuel injected is at least the maximum, pump torque control is effected by comparing the available torque with an amount of torque that is available at a combination of a set engine speed and an actual engine speed. A bias circuit biases the stable speed operating point a small amount below the maximum torque, thus producing a reference value, referred to as a predetermined value, whereby improved fuel economy is attained. In a system with more than one variable delivery pump, when less than the full number of pumps is employed, control is effected in the low-power range, even though all pumps in service are receiving maximum fuel. The present invention makes it possible, in the case where a rack command value exceeds the predetermined value, to adjust the required engine output torque T (corresponding to Point A in Fig. 3) which acts as a pump torque demand and corresponds to the set engine speed determined by means of an accelerating lever, to pumps and at the same time to further correct (i.e., increase) the pump torque demand so that the actual engine speed becomes lower than the set engine speed. Thus the engine operates in an area of more stability and the fuel consumption rate is improved. When the load on the engine is small, the rack command value is less than the predetermined value, this may be because only one pump is being driven. A relatively small engine output torque T will satisfy the pump torque demand. The pump torque demand is increased until the rack command value equals the maximum rack displacement, thereby increasing the load to a value which utilises nearly 100% of the engine power. The above, and other objects, features and advantages of the present invention will become apparent from the following description read in conjunction with the accompanying drawings, in which like reference numerals designate the same elements. Figure 1 is a block line drawing of a pump torque controller according to an embodiment of the present invention. Figure 2 is a schematic drawings of an engine pump controlling system including the above torque controller. Figure 3 is a graph showing the engine/torque characteristics and fuel consumption characteristics. Figure 4A-4B and Figure 5 are sets of graphs showing the pump power characteristics. Figure 6 is a schematic structure drawing of a conventional engine pump controlling system. In the description which follows, explanation of parts and elements that are identical to the prior-art control circuit of Fig. 6 is omitted. Briefly, the apparatus of the present invention employs a predictor of the amount of torque available at all commanded engine speeds as a reference for comparison with a processed error signal indicating the amount of torque actually being generated by the speed setting of the engine. A system using a selectable number of a plurality of pumps enables operation with less than full engine power. Referring to Figs. 1 and 2, engine 1 includes a fuel injection pump and a controller 7. Controller 7 controls the control rack (hereinafter referred to as the rack) of the fuel injection pump. A pump torque controller 8 serves controller 7. Actual engine speed 11 and rack command value 23 are conveyed from controller 7 to pump torque controller 8 via an electric signal line. An electric/oil pressure converter 9 produces a pump torque command for pump torque controller 8 via an electric signal line. The pump torque command is converted to an oil pressure command by swash plate regulators 4a and 4b. Fig. 1 illustrates the pump torque controller 8, and shows a set engine speed 10 determined by the setting of an accelerating lever or by other means and actual engine speed 11, which is the actual revolution rate of the engine. The pump absorption torque correction circuit operative in the lagging area (shown in Fig. 3) is comprised of underspeed control 12 of the engine speed from point A to point C in Fig. 3 (hereinafter referred to as under speed volume US), summers 13 and 14, proportional gain K2 15, integral gain K1 16, integral factor 17, summer 18 of proportional factors and integral factors, and conversion coefficient K3 19 to convert engine speed variation to a torque variation signal T F . The torque available from the engine at the selected speed is the engine output torque signal T E produced by circuit 20. The engine output torque signal T E is connected to the plus input of summer 21. The amount of torque that can be absorbed by the pumps is calculated and applied to the minus input of summer 21. The signal T E and the amount of torque that the pumps can absorb are subtracted from each other in summer 21 to produce the pump torque command When less than all of the pumps are used for absorbing engine torque, the swash plates are controlled in response to rack inputs and outputs, rather than speed inputs and outputs. In this situation, control is maintained in the governor area (Fig. 3). A pump absorption torque correction circuit which works in the governor area (shown in Fig. 3) comprises a preset maximum rack displacement 22, a rack command value 23, conveyed from rack controller 7 (the rack command value referred to herein is a command value to the engine rack regulators, and the present invention calls for utilizing this command value to control the pump swash plates), summer 24 for summing the maximum rack displacement 22 and the rack command value 23, integral gain K4 25, integral factor 26, summer 27, product K5 28, and summer 29 adding together for engine output torque T E and the output of proportional gain K5 28. The proportional gain provided by product K5 28 introduces a necessary conversion coefficient to convert rack variation to torque variation. Product K5 28 is hereinafter referred to as proportional gain K5. As shown in Fig. 1, a switch 30 switches between the pump absorption torque correction circuit working in the lagging area (Fig. 3) and that working in the governor area (Fig. 3). The "a" contact of switch 30 is used when the rack command value 23 is greater than or equal to the predetermined value (90 percent of maximum). In this circumstance, control responds to engine speed. The "b" side of switch is connected when rack command value 23 is less than the predetermined value. In this circumstance, control responds to rack commands and responses. Point A in Fig. 3 is a discontinuous intersecting point of the governor area and the lagging area. The unevenness of the curve at point A raises the possibility that stable operation may not be maintained in this vicinity. Therefore, in order to maintain stable operation, the operational condition is controlled to move beyond point A to point C. This control performed by the portion of pump torque controller 8 between actual engine speed 11 and conversion coefficient 19. Under speed volume (US) 12 applies a slight negative increment to set engine speed 10 so that the actual commanded speed applied to summer 14 is in the vicinity of point C in Fig 3. PI control is performed by a PI (proportion + integral) controller which includes the portions of pump torque controller 8 ranging from proportional gain 15 to summer 18 in order to make deviation delta N between the target speed and actual engine speed 11 approach zero. Since the output from the PI controller is in dimensions of the revolution rate, such output is converted to torque by the conversion coefficient 19. Integral factor 17 of the PI controller used in this stage has maximum and minimum values for improved control response. The output of the above conversion coefficient 19 is subtracted from the engine output torque T E at summer 21 to produce the pump torque command. Fig. 4A and Fig. 4B illustrate pressure versus discharge flow rate conditions for a system using two motors. Fig. 5 illustrates pressure versus discharge flow rate during one-motor operation. The horizontal axes of these figures indicate the discharge pressure of the pumps P1 and P2, and the vertical axes show the discharge rates of pumps Q1 and Q2. The curves (i.e., hyperbolas) shown as PS1 and PS2 in the figures can be used to calculate pump power (P1 X Q1 and P2 X Q2 respectively). The present invention calls for controlling the pump absorption torque (the load torque which the pump absorbs from the engine output torque in the form of pump discharge pressure multiplied by pump discharge flow rate). This is equivalent to controlling the operating positions along pump power curves PS1 and PS2 in the figures. Further, the power referred to hereinabove equals engine torque multiplied by engine speed. When the engine speed is constant, the ratio of engine power to engine torque is 1 : 1. With two pumps, the sum of power PS1 of one pump 2A and the power PS2 of the other pump 2B equals the minimum output of the engine. Referring to Fig. 1, in operation, when the actual engine speed 11 is lower than the target speed, the actual engine speed is found to the left of point C in Fig. 3. The engine is overloaded in this condition. The output delta N at summer 13 is positive, making the control output of the PI (proportion + integral) controller positive. This control output, modified by multiplication by positive conversion coefficient K3 to produce the signal T F , is subtracted from engine output torque T E at summer 21. When the pump torque command is decreased based on the above value, pump power moves along curves PS1 and PS2 toward the bottom-left portions of Figs. 4A and 4B. In short, in response to reduced engine loading, the engine speed is permitted to increase gradually. In contrast, if the actual engine speed 11 is greater than the target speed at point C in Fig. 3, the pump absorption torque is less than the target value. Delta N is negative, and the control output of the PI controller, times the conversion coefficient K3 19, applies a value T F to summer 21 that exceeds the engine output torque T E by an amount equal to conversion coefficient K3. This causes the operating points to move along the pump power curves PS1 and PS2 toward the right-upper portion of Figs. 4A and 4B. The resulting increased loading on the engine tends to lower the engine speed gradually until it is regulated in the vicinity of the target speed at point C. As illustrated in Fig. 3, it is evident that regulating the engine speed to the target speed point C by controlling the pump swash plates is advantageous with respect to fuel consumption efficiency. The following explanation is directed to operating when only a single pump controlled by the operation lever. Single-pump operation is used when the total load to the engine is small. Rack command value 23 from rack controller 7 is less than sthe predetermined value, and therefore contact "b" of switch 30 (Fig. 1) is selected. The control from maximum rack displacement 22 to summer 29 functions to make point B in Fig. 3 approach point A as closely as possible in the governor area. In this stage, too, engine output torque TE, described above, acts as the basis of control. Maximum rack displacement 22 corresponds to the rack displacement at point A in Fig. 3. Here, the output of PI controller, derived from integral gain 25 and fed to proportional gain 28, makes rack command value 23 identical to the maximum rack displacement 22. That is, it makes the difference delta R between them approach zero, thereby increasing the pump absorption torque. As described above, proportional gain 28 provides a factor which converts from the dimension of rack displacement to that of torque. The output of proportional gain 28 and engine output torque TE, which is the output of engine output torque characteristics 20, are added together at summer 29 to produce the pump torque command. The integral factor K4 25 in the PI controller has maximum and minimum values selected to prevent the engine from running away. The above condition is illustrated in the power characteristics curve of Fig. 5. The illustration uses the example wherein only pump 2A, producing power curve PS1, is effective. It will be noted that the position of the power curve PS1 in Fig. 5 is considerably higher and to the right of the corresponding curve in Fig. 4A. This difference indicates the increase in power required from a single pump. Nevertheless, if the maximum power of one pump is less than the maximum output of the engine, the operating point on the engine output torque characteristic curve of Fig. 3 cannot be moved far enough to reach point A. In that case, reversion to multiple-pump operation is indicated. Even when using two pumps, if the operating load is small, such as occurs when making fine adjustments, control similar to the single-pump control, described above, may be performed, with the pump torque command being output to the two pumps to produce a sum of powers approximating the fully loaded single-pump power shown in Fig. 5. Since swash plate regulators 4a and 4b are also driven by the pilot pressure of operational lever 5, the ability for fine adjustment is secured with implementation of an oil hydraulic pressure circuit which gives priority to the pilot pressure. As described as above, a control method according to the present invention enables the following: (1) in case of heavy load, where all the pumps are in operation, the engine output is utilized to its full extent in a safe area, that also operates the engine in an operating region good fuel efficiency can be achieved; and (2) in case of a lightly loaded engine using, for example, a single pump, the engine output is utilized to its full extent. With a pump torque control method according to the present invention, it is possible to utilize nearly 100% of the engine output to respond to a heavy load wherein all pumps are used. This operation is done with the engine output in a stable condition, thereby operating the engine with an efficient fuel consumption rate, by means of correcting the pump absorption torque, which is the basis of pump swash plate control, in such a manner that the actual engine speed is regulated to a value lower than the set engine speed. Under light load, nearly 100% of the engine power is used by correcting the pump absorption torque in such a manner that the rack command value is close to the maximum rack displacement.

Claims: An apparatus for controlling the torque in an engine and pump system wherein the engine (1) is coupled to drive a plurality of variable delivery pumps (2) the engine (1) having a fuel injection pump, comprising: an accelerating lever, a set engine speed device (10) responsive to the accelerating lever, an actual engine speed sensor (11), an engine output torque calculator (20) responsive to the set engine speed, characterised in that there is provided; a first correction circuit including; an under speed set means (12, 13), a first summer (14), a first proportional integral controller (15 to 18) to calculate a first correction torque (TF) according to the difference ( N) between a target engine speed and the actual engine speed, a second summer (21) to correct the engine output torque (TE) by the first correction torque (TF) in the lagging area, a second correction circuit including; a rack command value generator (23) responsive to the accelerator lever to generate a rack command value to control the fuel injection pump, a third summer (24), a second proportional-intregral controller (25 to 27) to calculate a second correction torque according to the difference ( R) between an actual rack command value and a maximum rack command value, a summer (29) to correct the engine output torque (TE) by the second correction torque in the governor area, a comparator which compares the rack command value with a predetermined rack command value related to the maximum rack command value to actuate a switch (30) whereby the pump is controlled by the first correction circuit when the rack command value is greater than or equal to the predetermined rack command value, and by the second correction circuit when the rack command value is less than the predetermined value.

Inventors: MORIYA NAOYUKI C O SHIN CATERP, SAMESHIMA MAKOTO C O SHIN CATE, MORIYA, NAOYUKI, SAMESHIMA, MAKOTO, MORIYA, NAOYUKI, C/O SHIN CATERPILLAR MITSUBISHI, SAMESHIMA, MAKOTO, C/O SHIN CATERPILLAR MITSUBISHI
Applicants: CATERPILLAR MITSUBISHI LTD, SHIN CATERPILLAR MITSUBISHI LTD.
Application year: 1990
Main classification: F04B1
IPC codes: F04B49, F02D29, F04B1