Abstract:
A hydraulic power steering system that maintains proportional pressure control between an input side of a gear and assist pressures internal to the steering gear by relying upon a control valve. A pressure differential is maintained between the input side of a gear and assist pressures internal to the steering gear throughout certain events. Those events include running the pump to fully charge the accumulator, idling the pump after fully charging the accumulator, discharging pressure from the fully charged accumulator at an onset of a steering event having a demand for steering load, recovering the discharged pressure of the accumulator by running the pump after turning on the pump at the onset of the steering event, and opening the relief valve in the valve manifold during the steering event.

Description:
BACKGROUND OF THE INVENTION 
   1. Field of the Invention 
   The invention relates to a hydraulic power steering system that continuously provides a source for hydraulic assist to a steering gear while allowing the pump to be idled when there is no steering demand to reduce energy consumption. 
   2. Description of Related Art 
   Conventional hydraulic power assist steering systems employ an open center valve at the steering gear and a pump, which supplies constant flow through an open loop hydraulic circuit. Fluid is pumped through the system at all times while the engine is running regardless of steering load. By closing ports within the steering valve creating pressure internal to the gear and in the hydraulic circuit upstream to the pump, boost or steering assist is created. Running a hydraulic pump continuously in this type of system ensures good response to a steering input but is inefficient when there is no steering load. Restrictions in the hydraulic system can create significant backpressure against the pump and generate a significant amount of heat over time that must be managed with oil coolers. As a result, prior studies have shown that fuel consumption in the unloaded mode can dominate the fuel consumed under load. 
   U.S. Pat. No. 5,921,342 (the &#39;342 patent) provides a Power Assisted Steering Apparatus for Automotive Vehicle, but does not have a proportional pressure relief for the pump under load but instead relies on throttling down flow through the system using a variable displacement pump when steering demand is low. Since the pressure drop across restrictions in the hydraulic system is proportional to flow rate, the backpressure that the pump needs to work against is lower when flow is reduced through the system in low steering demand situations. Therefore, this system reduces pumping losses and saves energy over hydraulic power steering systems that employ fixed displacement hydraulic pumps. 
   The patent application publication to Rogers et al. (2003/0127275 A1) provides for a High Efficiency Automotive Hydraulic Power Steering System and does have an automotive hydraulic power steering system that prevents wasted energy when no power assist is required but does not maintain proportional pressure control between the input side of the gear and the assist pressures internal to the gear. Proportional control is a key enabler to minimizing leakage through the steering valve and being able to control steering assist as a function of steering valve opening hydraulically without the need for electronic controls. The system described does not provide a load sensing signal downstream of the steering valve. Without a load sensing signal, the low pressure setting for the hysteresis pressure switch must be raised to ensure adequate response during worst case steering conditions which reduces the effective range of the accumulator. 
   SUMMARY OF THE INVENTION 
   One aspect of the invention resides in the design of a power steering hydraulic system that includes an engine driven power steering pump equipped with a clutch, a steering gear with a closed center control valve, a piston type accumulator, a valve control manifold and two pressure switches. The valve manifold provides pressure to the input side of the closed center steering valve that is proportional to the assist pressure internal to the gear so that the level of assist is both predictable and consistent as a function of the steering valve opening. A shuttle valve is used in conjunction with a pressure switch as the steering load sensing device. The accumulator is used as a storage device to permit the pump to be idled when there is no steering load and also to provide uninterrupted assist to the gear during pump clutch engagement at the onset of a steering event. 
   The inventor established the following design scheme in developing the invention with the objective of maximizing energy efficiency of hydraulic power steering systems—particularly those with higher steering load requirements. 
   Utilize the existing power source to drive the steering pump. Adding a separate power source (typically an electric motor) for power steering has clear advantages as it can be optimized for the application—Electric Power Steering (EPS) and Electro-Hydraulic Power Steering (EHPAS) are typical examples. However, the size and cost of electric motors needed for higher steering load applications make these systems undesirable or impractical at this point in time. 
   Turn off the pump when there is no steering load—As stated earlier, prior studies have shown that fuel consumption in the unloaded mode can dominate the fuel consumed under load. Therefore, this is a key to minimizing fuel consumption. 
   Use a closed center steering gear valve to eliminate fluid flow through the system when there is no steering load. Conventional open center valves require constant flow at all times. A closed center valve allows fluid to be stored upstream in the system and delivered to the gear on demand. 
   Add an accumulator to manage steering load/unload transitions. This enables initial steering assist (and/or brake assist if equipped with hydra boost) while the pump is being engaged at the onset of a steering (and/or braking) event. Prior studies with clutched pumps without accumulators have shown unacceptable response. The closed center steering gear valve provides flow on demand and in proportion to steering wheel rate rather than continuous flow, which minimizes accumulator size. 
   Use pressure internal to the gear and a switch as the steering load signal. Load sensing with a closed center steering valve can also be accomplished by sensing pressure drop across an orifice in the supply line to the gear or with a steering angle sensor on the gear input shaft. However, both load sensing signals have limitations. The pressure drop across an orifice can be difficult to measure when steering wheel rates (and consequently flow to the gear) are low but steering load and need for assist are high. Steering angle alone does not always correlate well to steering load and often requires other inputs, signal processing and calibration complications. 
   Use proportional hydraulic valve controls to control assist and pump flow. Proportional control allows supply and demand to balance hydraulically without the need for complicated and costly electronic control strategies. 

   
     BRIEF DESCRIPTION OF THE DRAWING 
     For a better understanding of the present invention, reference is made to the following description and accompanying drawings, while the scope of the invention is set forth in the appended claims. 
       FIG. 1  shows the hydraulic circuit diagram in accordance with an embodiment of the invention, at the start of the accumulator charge cycle with no steering load. 
       FIG. 2  shows the hydraulic circuit diagram in accordance with an embodiment of the invention, reflecting the end of the accumulator charge cycle with no steering load. 
       FIG. 3  shows the hydraulic circuit diagram in accordance with an embodiment of the invention, reflecting the onset of a light steering event. 
       FIG. 4  shows the hydraulic circuit diagram in accordance with an embodiment of the invention, reflecting the onset of a steering event that triggers pump clutch engagement. 
       FIG. 5  shows the hydraulic circuit diagram in accordance with an embodiment of the invention, reflecting accumulator fully charged during a steering event. 
       FIG. 6  shows the hydraulic circuit diagram in accordance with an embodiment of the invention, reflecting when the relief valve in the valve manifold begins to open during a steering event. 
       FIG. 7  shows the hydraulic circuit diagram in accordance with an embodiment of the invention, reflecting when the relief valve in the pump begins to open during a max load steering event. 
       FIG. 8  shows the hydraulic circuit diagram in accordance with an embodiment of the invention, reflecting release of the steering wheel after a max load steering event. 
       FIG. 9  shows the hydraulic circuit diagram in accordance with an embodiment of the invention, reflecting the pressure to the steering gear valve returning to set point through leakage at the steering gear valve. 
       FIG. 10  shows the hydraulic circuit diagram in accordance with an embodiment of the invention, reflecting accumulator pressure drop as flow is supplied to gear. 
       FIG. 11-13  show the electrical circuit diagram in accordance with the same embodiment of the invention and operation during the hydraulic power steering system events described in  FIGS. 1-10 . 
       FIG. 14-15  show the function of respective pressure switches as referenced in the hydraulic circuit diagrams shown in  FIGS. 1-10 . 
       FIG. 16  shows a free body diagram of the relief valve of  FIGS. 1-10 . 
       FIGS. 17-18  show the relief valve of  FIG. 16  in operation for accumulator charging, respectively, at a beginning of the accumulator charging cycle and at an end of the accumulator charging cycle. 
       FIGS. 19-20  show the relief valve of  FIG. 16  in operation for a steering event, respectively, closed and open. 
       FIG. 21  shows a response curve describing the function of control valve referenced in the hydraulic circuit diagrams shown in  FIGS. 1-10 . 
       FIG. 22  shows a response curve describing the function of relief valve referenced in the hydraulic circuit diagrams shown in  FIGS. 1-10 . 
   

   DETAILED DESCRIPTION OF THE INVENTION 
     FIGS. 1-10  share the following components in the hydraulic power steering system: a pump/clutch assembly  10 , an accumulator  12 , a brake hydra boost  14  with brake master cylinder  16 , a steering gear  19 , a solenoid valve  20  that is normally open, a relief valve  22  that is normally closed, a closed centered, (gear valve or) steering valve  18 , a shuttle valve  24 , a control valve  26  that is normally open, a check valve  28  downstream of the pump/clutch assembly  10 , a pressure switch  32  in the accumulator circuit, a pressure switch  34  in the reference pressure circuit downstream of shuttle valve  24  and reservoir  15 . 
   The clutch of the pump/clutch assembly  10  engages the pump when there is a steering demand or accumulator needs to be charged. Shuttle valve  24  monitors pressure on each side of the piston in steering gear  19  and provides the higher of the two pressures to the reference pressure circuit. Accumulator  12  is used to store power steering fluid under pressure. Flow from accumulator  12  allows control valve  26  to provide initial pressure/flow to steering gear  19  through steering valve  18  to prevent a response lag because of time delay from engagement of the pump/clutch  10  until the pump runs at applicable capacity during a steering event. 
   Solenoid valve  20  (normally open) allows relief valve  22  to close more rapidly when the clutch disengages after an accumulator charge. The solenoid in solenoid valve  20  energizes and closes the valve when the pump clutch is engaged blocking flow to reservoir  15  and allowing pump pressure to be sensed by relief valve  22 . The solenoid in solenoid valve  20  de-energizes and opens the valve when the pump clutch is disengaged allowing flow through solenoid valve  20  to reservoir  15 . This allows pressure to relief valve  22  to be vented to atmosphere driving the valve closed immediately after the pump clutch is disengaged since relief valve  22  is normally closed. 
   Relief valve  22  is normally closed and begins to open when pump supply pressure exceeds the pressure required to overcome the spring setting internal to relief valve  22  plus a secondary opposing pressure from the opposite end of relief valve  22  which is ported to the reference pressure circuit. A unique feature of relief valve  22  is that the area of the internal spool valve that is exposed to reference pressure is 30% smaller than the area of the spool valve that is exposed to the pump supply pressure. This configuration reduces the contribution of reference pressure on the spool and allows relief valve  22  to begin relieving some of the flow to the pump back to reservoir  15  as reference pressure increases to better balance pump supply with demand as determined by reference pressure. This valve design keeps the pump from going into relief prematurely in moderate to high steering load conditions. The operation of relief valve  22  is shown graphically in  FIG. 16 . 
   Schematic diagrams of the on-demand hydraulic power steering system in accordance with an embodiment of the invention are shown in  FIGS. 1-10 , each at different operating conditions. 
     FIG. 1  shows the hydraulic circuit diagram at the start of the accumulator charging cycle when there is no steering load (or brake load if the application includes hydra boost braking system). The pressure in accumulator  12  has been dropping due to small but expected leakage through steering valve  18 , which is currently closed and on center (i.e., no steering input by the driver). 
   In this example, pressure switch  32  is designed to close when pressure in the accumulator drops below the switch setting. In this example, switch  32  has just closed providing power to pump clutch  10  and solenoid valve  20 . Power to the normally open solenoid valve  20  closes the valve. As the pump clutch  10  engages, flow from the pump opens check valve  28  and begins to fill accumulator  12  while pressure builds in the accumulator circuit. Shuttle valve  24  is idle since there is no steering input from the driver and is supplying zero pressure to the reference pressure circuit. At this point, relief valve  22  is closed. 
   As pressure builds in the pump supply and accumulator circuits, relief valve  22  remains closed allowing all pump flow to charge accumulator  12 . The purpose of control valve  26  is to maintain proportional pressure control between the input side of the gear and reference pressure. In this example, control valve  26  is set to continuously maintain a pressure to the input side of the gear (ex. 300 psi) proportional to the valve spring setting. The operation of control valve  26  is shown graphically in  FIG. 21 . 
     FIG. 2  shows the hydraulic circuit diagram of the same embodiment reflecting the end of the accumulator charging cycle when there is no steering load (or brake load if application includes hydra boost). The pressure in accumulator  12  has been rising with the pump clutch  10  engaged. 
   In this example, pressure in the accumulator circuit has reached the maximum accumulator charge pressure and triggered the pressure switch  32  to open which turns off electrical power to pump clutch  10  and solenoid valve  20 . As pressure in the accumulator circuit approaches the trigger point of pressure switch  32 , relief valve  22  also limits pressure in the circuit during accumulator charging and begins to open when pump supply pressure reaches the maximum accumulator charging pressure (The reference pressure contribution in this case is zero since there is no steering load). 
   As pump clutch  10  disengages, flow from the pump stops, solenoid valve  20  opens, check valve  28  closes to prevent back flow and pressure immediately upstream of the check valve  28  is released to a vented reservoir  15 . Opening solenoid valve  20  allows pump supply reference pressure to relief valve  22  to drop and the valve to close immediately to prevent fluid and pressure loss from the accumulator circuit. Although pressure in the accumulator circuit is now at maximum charging pressure, control valve  26  continues to maintain a set pressure (ex. 300 psi) to the input side of the steering valve  18 , which is still closed and on center. Keeping the pressure low to steering valve  18  minimizes leakage through the valve, maximizes the time between accumulator charges and subsequently maximizes fuel efficiency. 
     FIG. 3  shows a hydraulic circuit diagram of the same embodiment reflecting when a light steering event begins. When the driver turns the steering wheel, the column shaft, the intermediate shaft between column and steering gear  19  turn as well (not shown). If the torque applied to the steering wheel is sufficient to overcome the torsion rate of a small steel bar (often referred to as a T-bar) hard mounted between the gear input shaft and valve sleeve, the steering valve  18  will begin to open. In  FIG. 3 , the light steering event is represented, i.e., by showing higher pressure on one side of steering gear  19  as steering valve  18  begins to open. 
   The pressure internal to the left side of the steering gear  19  has increased (ex. 150 psi) and that the ball in shuttle valve  24  has been forced to the right blocking flow to the other side of the gear. Shuttle valve  24  then feeds the full pressure internal to the left side of the steering gear to the reference pressure circuit. Since the piston inside gear  19  is tied directly to the vehicle road wheels through tie rods (not shown), the piston begins to move right in this case to turn the wheels. 
   As the piston moves to the right inside the gear housing, control valve  26  must provide flow through steering valve  18  to fill the increasing volume to the left of the gear piston and maintain proportional pressure control between the input side of the steering valve  18  and the reference pressure. The control valve  26  has increased supply pressure to the steering valve  18  (ex. 450 psi) to a level equal to the preset valve spring setting (ex. 300 psi) plus the reference pressure (ex. 150 psi). The accumulator  12  supplies flow passing through control valve  26  to the steering gear  19  and consequently the accumulator pressure drops (ex. 900 psi). The pressure in the accumulator circuit is between minimum and maximum accumulator charging levels so pressure switch  32  remains open so there is no power to the pump clutch. The reference pressure (ex. 150 psi) is below the trigger point for pressure switch  34  (ex. 300 psi), so the switch  34  remains open as well so the clutch to the pump is disengaged. 
     FIG. 4  shows the hydraulic circuit diagram of the same embodiment reflecting when a steering event triggers pump clutch engagement. As the driver continues a steering event (from  FIG. 2 ) and applies more torque to the steering wheel to steer the vehicle, steering valve  18  opens further allowing more flow and increases pressure to the left side of the steering gear  19 . Control valve  26  continues to maintain proportional pressure between the input side to steering valve  18  and reference pressure. Reference pressure is now above the trigger point for pressure switch  34  (ex. 300 psi), so the switch has just closed sending power to the pump clutch. Check valve  28  now opens as pump speed increases. 
   The control valve  26  continues to allow flow to the input side of steering valve  18  as the piston in the steering gear  19  moves right and has increased supply pressure (ex. 600 psi) proportionately above reference pressure (ex. 300 psi). Accumulator pressure has dropped (ex. 700 psi) and must have sufficient capacity to supply the input side of steering valve  18  while the pump clutch is engaging. Depending on component selection and system layout of the hydraulic system, pressure switch  32  and  34  settings may need to be adjusted to ensure adequate pressure and flow can be provided by control valve  26  to the input side of the steering valve  18  while the pump clutch is engaging during rapid steering inputs. 
     FIG. 5  shows the hydraulic circuit diagram of the same embodiment reflecting recovery of accumulator pressure during a steering event with pump clutch engaged. Since pumps are sized to deliver enough flow at engine idle to meet worst case steering requirements, output flow from the pump/clutch assembly  10  will exceed flow required at gear  19  under most conditions. The excess flow will then begin building pressure in the accumulator circuit. 
   In this case, pressure in accumulator  12  has fully recovered to the max accumulator pressure setting. Control valve  26  continues to increase pressure (ex. 1200 psi) to the input side of steering valve  18  in proportion to reference pressure (ex. 900 psi). The piston in gear  19  continues to move to the right as the driver continues to turn the steering wheel and steering load and assist pressures are increasing. Pressure switch  34  and solenoid valve  20  are closed. Relief valve  22  remains closed because the pump supply pressure (ex. 1300 psi) is still below the proportional cracking pressure of relief valve  22  (ex. 1570 psi=1300 psi spring+0.3×900 psi reference pressure). 
     FIG. 6  shows the hydraulic circuit diagram of the same embodiment reflecting the relief valve in the manifold beginning to open during a steering event. As the steering event continues, control valve  26  continues to increase pressure (ex. 1470 psi) to the input side of steering valve  18  in proportion to reference pressure (ex. 1170 psi). Also, as the pump continues to run, check valve  28  is open, and the pressure continues to rise in the accumulator circuit. Eventually pump supply pressure reaches the cracking pressure of relief valve  22 . 
   At this point, pump supply pressure (ex. 1600 psi) is nearly equal to the cracking pressure of relief valve  22  (ex. 1651 psi=1300 psi spring+0.3×1170 psi reference pressure). At this point, relief valve  22  begins to open allowing excess pump flow to return to reservoir  15 , which is vented to atmosphere. 
     FIG. 7  shows the hydraulic circuit diagram of the same embodiment reflecting the relief valve in the pump beginning to open during a maximum load steering event. Typically, maximum steering loads occur during a parking type maneuver when the steering wheel is turned all the way to the end of travel (steering stops). 
   In  FIG. 6 , relief valve  22  was beginning to open. However, as flow is directed to reservoir  15 , control valve  26  continues to increase pressure (ex. 1850 psi) to the input side of steering valve  18  in proportion to reference pressure (ex. 1850 psi) as steering loads approach maximum levels. All pumps have internal relief valves to protect the hydraulic system. In the example shown in  FIG. 7 , the relief valve internal to the pump is set to open at 1850 psi and it begins to open redirecting pump flow within the pump itself. As the flow drops, check valve  28  closes. At this point, relief valve  22  is also beginning to open, because the proportional cracking pressure of relief valve  22  (ex. 1855 psi=1300 psi spring+0.3×1850 psi reference pressure). Relief valve  22  provides a redundant fail safe to protect the system should the pump relief fail. 
     FIG. 8  shows the hydraulic circuit diagram of the same embodiment reflecting release of steering wheel (steering load to zero) after a maximum load steering event. When the steering wheel is released, steering valve  18  returns to the on-center position. On center, steering valve  18  is designed to allow flow between left and right side gear ports in gear  19  and reservoir  15  through a return port in the valve. As such, when the steering wheel is released pressure on each side of the piston in the steering gear  19  along with reference pressure through shuttle valve  24  are relieved to atmospheric pressure. This drives control valve  26  closed since it is designed to allow flow only when supply pressure to input side of steering valve  18  is less than reference pressure while the pressure differential at this point is much higher (ex. 1750 psi). Relief valve  22  closes since the pressure in the pump supply is zero and well below the cracking pressure of relief valve  22  (ex. 1300 psi=1300 psi spring+0.3×0 psi reference pressure). 
     FIG. 9  shows the hydraulic circuit diagram of the same embodiment showing pressure to steering valve  18  returning to the set point in control valve  26  (ex. 300 psi). This is accomplished by leakage through steering valve  18 , which is on center. 
     FIG. 10  shows the hydraulic circuit diagram of the same embodiment reflecting a steering event in the opposite direction. In this example, the steering event is light similar to the event described in  FIG. 2 . Reference pressure (ex. 150 psi) is still below the trigger point for pressure switch  34  so the electrical circuit to the pump/clutch  10  remains open. Note, that control valve  26  has opened again to provide flow to gear  19  and maintain proportional pressure control to steering valve  18 . 
     FIGS. 11-13  show three different modes for the electrical circuit for the same embodiment, which includes a battery that provides power to electrical components, pump/clutch assembly  10 , solenoid valve  20 , pressure switch  32 , and pressure switch  34 . 
     FIG. 11  reflects the standby mode where both pressure switch  32  and  34  are open. Since there is no path to ground for pump/clutch  10  and solenoid valve  20  there is no power to either component. The pump/clutch assembly  10  in  FIG. 11  is disengaged.  FIG. 12  reflects the accumulator charge mode where pressure switch  32  is closed and pressure switch  34  is open. Since the pressure switches are wired in parallel, there is a path to ground providing power to pump/clutch  10  and solenoid valve  20 . The pump/clutch assembly  10  in  FIG. 12  is engaged.  FIG. 13  reflects a steering event. The effect is the same as for  FIG. 12 , except that the pressure switch  32  is now open while switch  34  is closed. The pump/clutch assembly  10  in  FIG. 13  is engaged. 
     FIG. 14-15  show the function of the pressure switch  32  and the pressure switch  34  for the same embodiment.  FIG. 14  plots the function of switch position versus accumulator pressure for the pressure switch  32  while  FIG. 15  plots the function of switch position versus reference pressure for the pressure switch  34 . The hydraulic pressures at which the switches open and close vary depending on the application. The solid line arrows represent the response to increasing pressure, while the dashed line arrows represent the response to decreasing pressure. 
     FIGS. 16-20  show the relief valve  22  in various modes of operation.  FIG. 16  shows the relief valve  22  in a free body diagram. Directional arrows are depicted to show the force direction for the F(pump), F(reference) and F(spring). 
   The force balance equations for the relief valve  22  are:
 
 F (pump)= F (reference)+ F (spring)
 
Since Pressure=Force/Area,
 
 P (pump)* A 1 =P (reference)* A 2 +F (spring),
         with A 1  &amp; A 2  defining the cross sectional areas at each end of the spool valve according to the relation A 2 =A 1 *0.3 so that:
 
 P (pump)* A 1 =P (reference″ A 1*0.3 +F (spring).
   If the spring force is set so that
 
 F (spring)= P (accumulator fill setting)* A 1, then  P (pump)* A 1 =P (reference)* A 1*0.3 +P (accumulator fill setting)* A 1
   By dividing both sides by A1,
 
 P (pump)= P (reference)*0.3 +P (accumulator fill setting)
       

     FIGS. 17 and 18  depict operation of the relief valve  22  during the charging cycle of the accumulator  12  of  FIGS. 1-10 .  FIG. 17  represents the force balance on the relief spool during the beginning of the accumulator charging, while  FIG. 18  represents the force balance on the relief spool during the end of the accumulator charging. 
   The force balance equations for the relief valve  22 :
         Since the pump &amp; accumulator are in the same circuit,
 
 P (pump)= P (accumulator).
   If there is no steering event during the accumulator charging,
 
 P (reference)=0.
   Given the force balance equation:
 
 P (pump)= P (reference)*0.3 +P (accumulator fill setting),
   forces on the spool are balanced when P(pump)=P(accumulator fill setting). The valve will remain closed until P(pump)&gt;P(accumulator fill setting).       

     FIGS. 19 and 20  depict operation of the relief valve  22  during a steering event, with  FIG. 19  showing the relief valve  22  in a closed position and  FIG. 20  showing the relief valve  22  in an open position. When the pump clutch is engaged, the pump delivers constant flow to the system. Flow through a steering gear (with a closed center steering valve  18  of  FIGS. 1-10 ) varies based on the steering wheel rate and gear size. The function of the relief valve is to dump excess flow from the pump to a reservoir. The relief valve uses the pressure differential between the pump {P(pump)} and gear {P(reference)} to control flow. 
   For example, the force balance on the relief spool during a steering event:
         If P(accumulator fill setting)=1300 psi and P(reference) at a point in time=1170 psi, then the force balance on the relief spool is achieved when P(pump)=1651 psi. That is:
 
 P (pump)= P (reference)*0.3 +P (accumulator fill setting)=1170 psi*0.3+1300 psi=1651 psi
   Consequently: if P(pump) rises above 1651 psi, the relief valve  22  opens, but if the P(pump) falls below 1651 psi, the relief valve  22  closes.       

     FIG. 21  shows a graphical illustration of the function of the control valve for the same embodiment. The graph includes two lines: A reference line (solid line) showing when the supply pressure to the gear equals reference pressure and a dashed line that represents the pressure provided by the control valve to the gear as a function of reference pressure. The dashed line describing the control valve function also shows the condition when the pump goes into relief. Increasing the valve pressure setting above the gear reference pressure increases leakage through the steering valve, decreases the time between successive accumulator cycles, may improve system response and/or reduce catch, and forces the pump into relief at lower steering load conditions. 
     FIG. 22  shows a graphical illustration of the function of the relief valve for the same embodiment. The graph includes two lines: A reference line (solid line) showing when accumulator pressure equals reference pressure and a dashed line that represents the pressure provided by the relief valve on the accumulator circuit as a function of reference pressure. Decreasing the area of the valve spool exposed to the reference pressure relative to the spool area exposed to the accumulator pressure flattens the slope of the line and prevents the pump from going into relief too early, allows higher accumulator storage pressure settings than would otherwise be the case, increases the time between accumulator cycles when there is no steering demand, reduces energy consumption. 
   While the foregoing description and drawings represent the preferred embodiments of the present invention, various changes and modifications may be made without departing from the scope of the present invention.