Patent Publication Number: US-6209331-B1

Title: Air handling controller for HVAC system for electric vehicles

Description:
CROSS REFERENCE TO RELATED APPLICATIONS 
     This application is a division of U.S. Ser. No. 09/190,473, filed Nov. 12, 1998, U.S. Pat. No. 6,077,158. 
    
    
     BACKGROUND AND SUMMARY OF THE INVENTION 
     The present invention relates generally to automotive HVAC systems for controlling the environment of an automobile passenger compartment. More particularly, the invention relates to an air handling system for controlling the positioning of the fresh/recirc door in an automotive HVAC system. 
     This application is related to co-pending applications titled Refrigerant Flow Management Center For Automobiles, Reversible Air Conditioning And Heat Pump HVAC System For Electric Vehicles, Controller For Reversible Air Conditioning And Heat Pump HVAC System For Electric Vehicles, Anti-Fog Controller For Reversible Air Conditioning And Heat Pump HVAC System For Electric Vehicles, Controller For Heating In Reversible Air Conditioning And Heat Pump HVAC System For Electric Vehicles, and System For Cooling Electric Vehicle Batteries. Each of these applications is incorporated by reference into the present application. 
     Automotive heating ventilation and air conditioning, HVAC, systems have traditionally been single loop designs in which the full volume of refrigerant flows through each component in the system. In an HVAC system, refrigerant in the vapor phase is pressurized by a compressor or pump. The pressurized refrigerant flows through a condenser which is typically configured as a long serpentine coil. As refrigerant flows through the condenser heat energy stored in the refrigerant is radiated to the external environment resulting in the refrigerant transitioning to a liquid phase. The liquefied refrigerant flows from the condenser to an expansion valve located prior to an evaporator. As the liquid flows through the expansion valve it is converted from a high pressure, high temperature liquid to a low pressure, low temperature spray allowing it to absorb heat. The refrigerant flows through the evaporator absorbing heat from the air that is blown through the evaporator fins. When a sufficient amount of heat is absorbed the refrigerant transitions to the vapor phase. Any further heat that is absorbed raises the vaporized refrigerant into the superheated temperature range where the temperature of the refrigerant increases beyond the saturation temperature. The superheated refrigerant flows from the outlet of the evaporator to the compressor where the cycle repeats. Generally, the refrigerant flowing into the compressor should be in the vapor phase to maximize pumping efficiency. The operation of the refrigerant loop in conventional automotive HVAC systems is controlled by cycling the compressor on and off, and by varying the volume of refrigerant that is permitted to flow through the expansion valve. Increasing the volume of refrigerant that flows through the valve lengthens the distance traversed by the liquid before it changes to the vapor phase, allowing the heat exchanger to operate at maximum efficiency. 
     Advances in automotive HVAC systems have led to zone temperature control systems wherein different zones of an automobile are independently controlled. Zone control systems generally include an evaporator and expansion valve for each zone. The refrigerant flows through a compressor and condenser, then is split by a system of valves before flowing to the expansion valve and evaporator of each zone. The refrigerant flowing out of the evaporator of each zone is then recombined before returning to the compressor. 
     Further advances in automotive HVAC systems has led to the implementation of reversible heat pump systems in automobiles. In a reversible heat pump system the HVAC system can either heat or cool a compartment depending on the direction of the refrigerant flow. In the air conditioning mode refrigerant flows from the compressor through an outside coil (condenser) and into an expansion valve and inside coil (evaporator) before returning to the compressor. Heat energy is extracted from air that is blown through the inside coil (evaporator) into the passenger compartment thus providing cooled air. In the heating mode a four way valve reverses the flow of refrigerant through the coils, thereby reversing the function of the coils. Refrigerant flows from the compressor through the inside coil (condenser) then into an expansion valve and the outside coil (evaporator) before returning to the compressor. Heat energy in the liquefied refrigerant flowing through the inside coil is absorbed by air that is blown through the coil into the passenger compartment thus providing heated air. The air that is blown through the coil is a mixture of fresh outside air and air that is recirculated from the passenger compartment. 
     Generally, in conventional systems the precise mixture of fresh air to recirculated air is selected by the vehicle occupants. Permitting passengers to exercise absolute control over the air mixture selection normally enhances the comfort of the passengers. However, under some operating conditions it leads to reduced passenger comfort and less than optimal vehicle performance. For example, when a reversible HVAC system switches from air conditioning mode to heat pump mode fogging of the vehicle windows that occurs from moisture evaporating into the conditioned air is more likely if only recirculated air is flowing into the passenger compartment. Another example is air blow-by which occurs when a vehicle increases beyond a particular speed. When an intermediate air mixture setting is selected the increased air pressure from the speed of the vehicle causes air to flow back through the recirc ducts and out the inlets. The air exiting the inlets is unconditioned, directly subjecting passengers to outside air. Not providing automatic override of the air mixture setting can subject passengers to degraded operating conditions in which the solutions are not obvious to the passenger. 
     One object of the present invention is to provide a system for selectively overriding the passenger air mixture selection to enhance passenger comfort. 
     Another object of the present invention is to improve vehicle performance by automatically adjusting the air mixture during predetermined vehicle operating modes. 
     A further object of the present invention is to reduce the heat load on the HVAC system. 
     Accordingly, the invention provides an air-flow management system for controlling the supply air to a motor vehicle passenger compartment. The air-flow management system includes a reversible heat pump system for transferring heat energy between an outside environment and a refrigerant. Air from the outside environment, fresh air, and from the passenger compartment, recirculated air, is forced through the air-flow structure by a blower resulting in the transfer of heat energy between the refrigerant and the passenger compartment. A recirculation door provides a means for controlling the mixture of fresh air to recirculated air that flows through the air-flow structure. The position of the recirculation door is selectable by a controller to prevent fogging during the transition from cooling mode to heating mode, minimize the energy expended conditioning the passenger compartment air, and prevent the backflow of unconditioned outside air from the fresh air duct into the recirculation duct. 
     The above described device is only an example. Devices in accordance with the present invention may be implemented in a variety of ways. 
    
    
     BRIEF DESCRIPTION OF THE DRAWINGS 
     These and other objects of the present invention will become apparent to those skilled in the art from the following detailed description in conjunction with the attached drawings in which: 
     FIG. 1 is a schematic representation of a preferred embodiment of the automotive refrigerant circuit; 
     FIG. 2 is a cross-sectional view of the flow management center shown in FIG. 1; 
     FIGS. 3 a  and  3   b  present cross-sectional views of flow management devices embodying the present invention; 
     FIG. 4 is a schematic representation of an alternative automotive refrigerant circuit; 
     FIG. 5 is a block diagram illustration of the control circuit interconnection to a reversible HVAC refrigerant circuit; 
     FIG. 6 is a flow diagram showing an overview of the control program for the preferred embodiment of the invention; 
     FIG. 7 is a flow diagram illusion of the expansion valve control program for the preferred embodiment of the invention; 
     FIG. 8 is a flow diagram of the compressor speed control module for the preferred embodiment of the invention; 
     FIG. 9 is a diagram illustrating the interaction between the expansion valve and compressor during the turn-on transition; 
     FIG. 10 is a datagram illustrating the relationship between the temperature cycle and a schematic representation of an HVAC system; 
     FIG. 11 is a flow diagram of the anti-fog algorithm for the preferred embodiment of the invention; 
     FIG. 12 is a flow diagram of the heating mode selection module for the preferred embodiment of the invention; 
     FIG. 13 is a flow diagram of the air-handling method for the preferred embodiment of the invention; 
     FIG. 14 is a schematic representation of a preferred embodiment of an HVAC system coupled to a battery pack module; 
     FIG. 15 is datagram illustrating the relationship between a preferred embodiment of the HVAC system and its heat load cycle; and 
     FIG. 16 is a datagram illustrating the relationship between the temperature lever position and the corresponding operating mode. 
    
    
     DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENT 
     FIG. 1 illustrates an exemplary reversible HVAC system  50  for motor vehicles that includes an air-flow structure  52 , a refrigerant flow system  54 , and a front panel  55  for providing controlling inputs. The reversible HVAC system  50  can both heat and cool the passenger compartment air of a motor vehicle by using the refrigerant flow system  54  in conjunction with the air-flow structure  52  to transfer heat energy between the outside environment and the passenger compartment. In heating mode, heat energy is transferred from the outside environment to air that flows into the passenger compartment and in cooling mode, heat energy is transferred to the outside environment from air that flows into the passenger compartment. The refrigerant flow system  54  acts as a storage medium for heat energy that is being transferred between the outside environment and the passenger compartment. The air-flow structure  52  controls the flow of conditioned air into the passenger compartment. An inside heat exchanger  88  provides an interface between the refrigerant flow system  54  and the air-flow structure permitting the transfer of heat energy between the refrigerant and the air flowing into the passenger compartment. The front panel  55  provides a means for the passengers to control the temperature, flow rate, and operating mode of the HVAC system. 
     The air-flow structure  52  includes a duct  56  through which air is supplied into the passenger compartment, a blower  58  for introducing air into the duct  56 , a recirculation door  60  for controlling the proportion of fresh air to recirculated air, a PTC heater  62  for heating the air, a blend door  60  for controlling the proportion of air that flows over the PTC heater  62 , and a set of duct outlets for discharging air into the passenger compartment. 
     The duct outlets include a defrost outlet  64  for directing air towards the windshield of the vehicle, a panel outlet  66  for directing air towards the upper extremities of the passengers, and a floor outlet  68  for discharging air towards the lower extremities of the passengers. The duct outlets  64 - 68  are selectively opened and closed by a mode damper  70  which operates in accordance with the position of the mode selector switch  72  located on front panel  55 . 
     The refrigerant flow system  54  is operable in a heating mode and a cooling mode and includes a compressor  76 , a four-way switch  78  for controlling the direction of refrigerant flow, an inside heat exchanger  88  for transferring energy between the refrigerant and air flowing into the passenger compartment, an outside heat exchanger  80  for interfacing with the outside environment, a flow management center  82  for reducing the pressure of refrigerant flowing into a heat exchanger that is functioning as an evaporator, shut-off valves  84  and  86  for system protection, zone-control heat exchanger  92  for providing independently controlled cooling to a local region, and pressure reducing device  90  for reducing the pressure of refrigerant flowing into the zone-control heat exchanger  92 . The refrigerant flow system  54  interacts with the air-flow structure  52  and the passenger compartment through the operation of the inside heat exchanger  88  during the heating and cooling modes. The function of the inside heat exchanger  88  changes in each operating mode; during heating mode the inside heat exchanger  88  functions as a condenser transferring heat energy to air that passes through air-flow structure  52  into the passenger compartment and during cooling mode the inside heat exchanger  88  functions as an evaporator absorbing heat energy from the air that passes through air-flow structure  52  into the passenger compartment. 
     The compressor  76  is driven by a variable speed electric motor (not shown). Varying the speed of the electric motor causes a commensurate change in the suction pressure and refrigerant discharge capacity of compressor  76 . Although the compressor in the present embodiment is a variable speed compressor, it is within the scope of the invention to employ a single speed compressor. The four-way switch  78  is connected between the compressor  76  and the heat exchangers  80  and  88  to provide a method of changing from air conditioning mode to heat pump mode by reversing the direction of refrigerant flow. 
     The inside heat exchanger  88  functions as an evaporator during a cooling operation and as a condenser during a heating operation. Inside heat exchanger  88  is arranged within duct  56  so that the air blown through the exchanger  88  is conditioned prior to passing over PTC heater  62  and being discharged through the duct outlets. Shut-off valve  84  provides a means of interrupting refrigerant flow during HVAC operating modes that do not require operation of inside heat exchanger  88 . Examples of such operating modes include disabling operation of the inside heat exchanger  88  as an evaporator at low ambient temperatures that could result in freezing of the heat exchanger  88  due to condensation, and modes where only secondary heat exchangers are operational such as zone control heat exchanger  92 . Such operating modes include cooling of a battery assembly and cooling of pre-selected regions within the vehicle. The flow management center  82  reduces the pressure of and expands the refrigerant to be supplied to the inside heat exchanger  88  during a cooling operation. 
     The outside heat exchanger  80 , which is generally located towards the front of the vehicle, exchanges heat between the outside air and the refrigerant. A fan  94  ensures a constant supply of outside air flows through outside heat exchanger  80 . During air conditioning mode the outside heat exchanger  80  functions as a condenser providing a means for the refrigerant to shed heat to the outside air. During heat pump mode the outside heat exchanger  80  functions as an evaporator absorbing heat energy from the outside air into the refrigerant. 
     The flow management center  82  provides a centrallized device for reducing the pressure of refrigerant flowing into a heat exchanger  80  or  88  functioning as an evaporator and acts as a source of high pressure liquid refrigerant for secondary heat exchangers. Conventional circuits use a separate pressure reducing device with bypass plumbing for each heat exchanger that functions as an evaporator. By using a single flow management center  82  to provide pressure reduced refrigerant the complexity of the refrigerant flow system  54  is greatly reduced. Additionally, a receiver/drier function is integrated into the flow management center  82  for eliminating contaminants and providing a reservoir of pressurized liquid refrigerant. Refrigerant tapped from the receiver portion is routed to pressure reducing device  90  and then to zone-control heat exchanger  92 . Although the flow management center in the preferred embodiment includes a receiver/drier function the principles of the invention can be extended to flow management devices that do not include a receiver/drier function. 
     Flow management center  82  is illustrated in greater detail in FIG. 2 to include a housing  100  defining bi-directional ports  102  and  104 , a pressure sensitive valve  106 , check valves  108  and  110 , desiccant  112 , a uni-directional flow member  114 , pressure reducing valve  116 , outlets  118  and  120 , temperature probe  124 , and pressure probe  122 . Pressurized liquid refrigerant flows into bi-directional port  102  or  104 , through the corresponding check valve  108  or  110 , through the dessicant  112 , into reservoir  113 , up the unidirectional flow member  114 , through pressure reducing device  116  and pressure sensitive valve  106 , and finally reduced pressure refrigerant flows out of the other bi-directional port  104  or  102 . When the HVAC system  50  changes operating modes the direction of refrigerant flow reverses as high pressure refrigerant flows into the bi-directional port that pressure reduced refrigerant was flowing from. The refrigerant then flows through the corresponding check valve  110  or  108 , through the dessicant  112 , into reservoir  113 , up the uni-directional flow member  114 , through pressure reducing device  116  and pressure sensitive valve  106 , and finally reduced pressure refrigerant flows out of the other bi-directional port  102  or  104 . The pressure sensitive valve  106  permits the flow of pressure reduced refrigerant out of one bi-directional port while preventing high pressure refrigerant from flowing directly between the bi-directional ports. When high pressure refrigerant flows into a bi-directional port  102  and  104  the pressure sensitive valve  106  closes the flow path from the port to the pressure reducing device  116  and opens a path from the pressure reducing device to the other bi-directional port  104  and  102 . Closing the flow path from the bi-directional port  102  or  104  to the pressure reducing device forces refrigerant to flow through the corresponding check valve  108  or  110 , through the dessicant  112 , and into reservoir  113 . The opposing check valve  110  or  108  prevents high pressure liquid refrigerant in reservoir  113  from flowing out the opposing bi-directional port  104  or  102 . Impurities within the refrigerant are removed by dessicant  112 . Reservoir  113  provides a pool of high pressure liquid refrigerant that can be sourced to multiple pressure reducing devices such as device  116  within the flow management center  82  as well as pressure reducing devices that provide reduced pressure refrigerant to secondary heat exchangers. Outlets  118  and  120  provide a means of tapping off refrigerant from reservoir  113  and directing it to secondary heat exchanger circuits. In the preferred embodiment the pressure sensitive valve  106  is a dual poppet valve, however it is envisioned that other valves such as multiple check valves, mushroom valves, reed valves, or rotary valves may be employed. Additionally, similar valves as listed above can replace check valves  108  and  110 . Although the pressure reducing device  116  in the preferred embodiment is an electronically controlled expansion valve it is within the scope of the invention to use mechanically controlled expansion valves as well as 90° valves. The desiccant  112  and the temperature and pressure probes  122  and  124  are merely exemplary of additional functions that can be added to the flow management center, they are not required to practice the invention. 
     Returning to FIG. 1, the zone-control heat exchanger  92 , located within the interior of the vehicle provides cooling functions for local zones or assemblies. Examples of local zone cooling include battery assembly cooling, air conditioned seats, and individualized cooling of one side of the passenger compartment. Pressure reducing device  90  reduces the pressure of and expands the refrigerant to be supplied to zone control heat exchanger  92 . The expanded refrigerant absorbs heat from the air or liquid which is passed through heat exchanger  92 , thereby cooling the air or liquid. 
     The front panel  55  includes selector switches for setting the operating parameters of the air conditioning circuit  50 . The switches include a blower speed selector  73  that in the preferred embodiment is adjustable from 30% to 100% of the maximum blower speed, a mode selector switch  72  having five mode settings, a recirculation selector  75  for selecting fresh or recirculated air, and a sliding temperature lever  74  for setting the temperature of air discharged from the duct outlets. Although the mode selector switch in the preferred embodiment has five discrete settings, the principles of the invention can be extended to a mode selector having an unlimited number of settings. 
     During cooling mode, the refrigerant discharged from the compressor  76  flows through four-way switch  78  into outside heat exchanger  80  which functions as a condenser. As heat energy stored in the refrigerant is shed to the outside air which is blown through the exchanger  80  the refrigerant condenses to a high pressure liquid. The liquid refrigerant flows into a bi-directional port  102  of the flow management center  82 , through the desiccant  112 , into the reservoir  113 , up the uni-directional flow member  114 , through the pressure reducing valve  116 , and then out the other bi-directional port  104 . A portion of the refrigerant is tapped off from the reservoir  113  and directed towards a secondary loop as shown in FIG. 1 will be explained in a later paragraph. The refrigerant flowing through the pressure reducing valve  116  is pressure reduced and then passes through the other bi-directional port  104 . The pressure reduced refrigerant flows into the inside heat exchanger  88  which functions as an evaporator. Heat energy from air passing through inside heat exchanger  88  is absorbed by the pressure reduced refrigerant causing the refrigerant to change to the vapor state. The vapor state refrigerant flows from the heat exchanger  88  through the four-way switch  78  and back to the inlet of compressor  76  which compresses the vapor and directs it through four-way switch  78  to outside heat exchanger  80 . 
     The operation of the secondary loop during cooling mode is as follows. The portion of refrigerant that flowed from an outlet in reservoir  113  flows through shut-off valve  86  into pressure reducing device  90 . Pressure reduced refrigerant flows out of device  90  into local-zone heat exchanger  92  which functions as an evaporator. The refrigerant absorbs heat from the air which passes through it thereby providing separately controlled cooling for a portion of the passenger compartment. Although the zone control heat exchanger  92  in the preferred embodiment functions as an air-to-refrigerant evaporator, it is within the scope of the invention to employ other heat exchangers such as refrigerant-to-refrigerant, water-to-refrigerant, and oil-to-refrigerant heat exchangers. 
     During heating mode, the direction of refrigerant flow is reversed by changing the orientation of four-way switch  78 . A signal from a controller  130 , hereinafter described, controls the orientation of four-way switch  78 . The refrigerant discharged from the compressor  76  flows through four-way switch  78  into inside heat exchanger  88  which functions as a condenser. As heat energy stored in the refrigerant is shed to the inside air which is blown through the exchanger  88  the refrigerant condenses to a high pressure liquid. The liquid refrigerant flows into the bi-directional port  104  of the flow management center  82 , through the desiccant  112 , into the reservoir  113 , up the unidirectional flow member  114 , through the pressure reducing valve  116 , and then out the other bi-directional port  102 . The refrigerant flowing through the pressure reducing valve  116  is pressure reduced and then passes through bi-directional port  102 . The pressure reduced refrigerant flows into the outside heat exchanger  80  which functions as an evaporator. Heat energy from air passing through outside heat exchanger  80  is absorbed by the pressure reduced refrigerant causing the refrigerant to change to the vapor state. The vapor state refrigerant flows from the heat exchanger  80  through the four-way switch  78  and back to the inlet of compressor  76  which compresses the vapor and directs it back through four-way switch  78  to inside heat exchanger  88 . 
     During heating mode, the secondary loop operates in the same manner as during a cooling mode. Refrigerant from outlet  118  of flow management center  82  flows through pressure reducing device  90  and into local-zone heat exchanger  92  in which it absorbs heat from air that is passing through the exchanger  92 . Pressure reducing device  90  pressure reduces the refrigerant to increase its capacity to absorb heat energy from air or fluid flowing through the heat exchanger  92 . 
     Employing flow management center  82  in reversible HVAC system  50  greatly simplifies the interconnecting plumbing and permits more reliable implementation of secondary cooling loops. It is possible to alternately heat and cool a vehicle with two heat exchangers without the additional valves and plumbing required for conventional systems. Complex refrigerant balancing schemes for dividing refrigerant amongst multiple heat exchanger loops are not required, thereby improving system performance, increasing system reliability, and reducing cost. A common sense point at the outlet of pressure reducing device  116  is provided for pressure reduced (low-side) refrigerant. Sensing temperature and pressure at the flow management center eliminates the need of conventional systems for sensing at the inlet to each heat exchanger. 
     Referring to FIGS. 3 a  and  3   b,  an alternate flow management device  81  is illustrated which does not include the receiver/drier function, but provides reversibility with simpler plumbing than conventional systems and a single pressure reducing device. The flow management device includes a housing  100  defining bi-directional ports  102  and  104 , a pressure sensitive valve  106 , check valves  108  and  110 , a unidirectional flow member  114 , temperature probe  122 , and pressure probe  124 . The flow management device  81  includes all the capabilities of the flow management center  82  with the exception of the receiver/drier function. 
     FIG. 4 illustrates another embodiment of an automotive air conditioning circuit  40  that includes a compressor  41 , an outside heat exchanger  42 , an inside heat exchanger  43 , two four-way switches  44  and  45 , a receiver/drier  46 , and an electronic expansion valve  47 . 
     Four-way switch  45 , receiver/drier  46 , and expansion valve  47  functionally replace the flow management center  82  that is employed in circuit  50  (see FIG.  1 ). The function of four-way valve  45  is the mirror image of the function of four-way valve  44 . Valve  44  is employed to reverse the flow of refrigerant through the heat exchangers  42  and  43 . It essentially converts unidirectional refrigerant flow from the compressor  41  into bi-directional refrigerant flow into the heat exchangers  42  and  43 . Whereas four-way valve  45  converts bi-directional refrigerant flow from the heat exchangers  42  and  43  into a unidirectional flow through receiver/drier  46  and expansion valve  47 . 
     Receiver/drier  46  removes contaminants from the refrigerant and ensures a continuous flow of high pressure liquid refrigerant into expansion valve  47 . Expansion valve  47  provides refrigerant pressure reduction and expansion for heat exchangers  42  and  43 . Expansion valve  47  is preferably an electronic expansion valve that receives its controlling inputs from a controller that monitors the saturation and superheat temperature of the heat exchangers  42  and  43 . However, other pressure reducing devices such as block valves, 90° valves, and thermal expansion valves (TXV) are within the scope of the invention. Generally, to control a TXV, refrigerant at the superheat temperature and the saturation temperature must be routed to the device. To obtain the superheat temperature the refrigerant from four-way valve  44  to the compressor  41  inlet can be routed through the TXV. For the saturation temperature the refrigerant emitted from the TXV can be sensed. 
     During cooling mode outside heat exchanger  42  functions as a condenser shedding heat to the outside environment and inside heat exchanger  43  functions as an evaporator absorbing heat from air that is blown into the passenger compartment. The refrigerant cycle is as follows: refrigerant flows out of compressor  41 , through four-way valve  44 , into the outside heat exchanger  42 , through four-way valve  45 , into receiver/drier  46  and expansion valve  47 , through four-way valve  45 , to inside heat exchanger  43 , through four-way valve  44 , and back to compressor  41 . 
     During heating mode four-way valve  44  changes orientation causing the flow of refrigerant to heat exchangers  42  and  43  to reverse. With the reversal in the direction of refrigerant flow the functions of the heat exchangers  42  and  43  reverse as inside heat exchanger  43  functions as a condenser and outside heat exchanger  42  functions as an evaporator. In addition, the orientation of four-way valve  45  is also changed to ensure that the direction of refrigerant flowing into receiver/drier  46  and expansion valve  47  remains constant. The refrigerant cycle during heat pump mode is as follows: refrigerant flows out of compressor  41 , through four-way valve  44 , into the inside heat exchanger  43 , through four-way valve  45 , into receiver/drier  46  and expansion valve  47 , through four-way valve  45 , to outside heat exchanger  42 , through four-way valve  44 , and back to compressor  41 . 
     From the foregoing it will be understood that the invention provides a flow management device with bi-directional ports in which refrigerant flowing into either port passes through an expansion valve and exits the other port. Additionally, the invention can integrate the receiver/drier function into a flow management device with bi-directional ports to provide the capability of tapping off refrigerant flow for secondary cooling circuits. Also, the present invention decreases the complexity of automotive HVAC systems by integrating a flow management device into the system to reduce the number of valves required to implement a reversible heating and cooling HVAC system. A further capability of the invention is to provide a centralized flow management center with taps for refrigerant to reduce the complexity of automotive HVAC systems that implement multi-zone control. 
     Control System For Reversible Air Conditioning And Heat Pump HVAC System For Electric Vehicles 
     FIG. 5 illustrates the control system configuration to implement the preferred embodiment of the HVAC circuit  50 . In FIG. 5 the outside coil  80 , flow management center  82 , inside heat exchanger  88 , four-way switch  78 , compressor  76 , duct  56 , and front panel  55  are interconnected in a manner similar to circuit  50  illustrated in FIG.  1 . Additionally illustrated is controller  130  which controls the compressor speed and flow management center  82  operation based upon inputs from front panel  55 , duct  56 , and the refrigerant system  54 . 
     During operation of the HVAC circuit  50 , the passenger selects a passenger compartment temperature and operating mode by setting the switches of front panel  55 . The front panel  55  switch settings are decoded by the controller  130 , which converts the settings to values that represent desired temperature, operating mode, and blower speed. The controller  130  also monitors sensors that measure the actual ambient and passenger compartment temperature as well as refrigerant temperature and pressure. The controller  130  compares the decoded settings to the actual ambient and passenger compartment temperature, and generates signals that modify the operation of the refrigerant flow system  54  and air-flow structure  52  to bring the actual passenger compartment temperature in conformance with the desired temperature as represented by the front panel  55  switch settings. 
     The operation of the refrigerant flow system  54  is modified by controller  130  through output signals that control the orientation of the four-way switch  78 , the speed of compressor  76  and the duty cycle applied to the pressure reducing device  116  within the flow management center  82 . Changing the orientation of four-way switch  78  causes a reversal in the direction of refrigerant flow. The direction that refrigerant flows dictates whether the HVAC system is in the heating mode or the cooling mode by interchanging the functions of the outside heat exchanger  80  and the inside heat exchanger  88 . In heating mode the outside heat exchanger  80  functions as an evaporator and the inside heat exchanger  88  functions as a condenser  88 . Whereas, in cooling mode the outside heat exchanger  80  functions as a condenser and the inside heat exchanger  88  functions as an evaporator. Varying the speed of compressor  76  during a cooling mode or a heating mode causes a change in the refrigerant temperature at the compressor  76  inlet and outlet, which has a direct effect on the temperature of air blown into the passenger compartment. Changing the duty cycle applied to the pressure reducing device  116  during either cooling or heating mode causes a variation in the quantity of refrigerant that the pressure reducing device  116  permits to flow into the heat exchanger  80  or  88  that is functioning as an evaporator. Too much refrigerant flowing through the evaporator leads to flooding the compressor  76 , causing degraded compressor  76  performance. Too little refrigerant flowing through the evaporator limits the efficiency of the evaporator in absorbing heat, resulting in a reduced cooling or heating capacity of the HVAC system  50 . The controller  130  constantly adjusts the duty cycle applied to the pressure reducing device to keep the evaporator operating at maximum efficiency and adjusts the speed of compressor  76  to control the temperature of the air blown into the passenger compartment. 
     The air-flow structure  52  operation is modified by changing the position of blend door  61  and the position of recirculation door  60 . Changing the position of blend door  61  changes the amount of supplemental electric heating that is applied to the air flowing through the air-flow structure  52 , directly effecting the temperature of the passenger compartment. The position of recirculation door  60  controls whether fresh air from the outside or recirculated air from inside is directed into the passenger compartment. Typically, more energy is required to heat or cool fresh air than recirculated air because of the greater differential between the temperature of the air flowing into the HVAC system  50  and the desired passenger compartment temperature. 
     Inputs to controller  130  from the front panel  55  include blower speed from blower speed selector  73 , mode selection from mode selector switch  72 , and the target temperature from temperature lever  74 . The duct  56  inputs include inlet and outlet temperatures from temperature probes  132 ,  133 , and  134 . Inputs from the refrigerant system  54  to the controller  130  include temperature probe  135  for sensing ambient temperature, temperature probe  124  for sensing the expansion valve  116  outlet temperature, temperature probe  136  for sensing superheat temperature, and pressure probe  138  for sensing suction pressure. 
     Controller  130  is preferably a microprocessor-based controller, that includes a processor  140  and associated memory  142 . An analog-to-digital converter (A/D)  144  converts signals from the various sensors to a digital form used by processor  140 . A driver circuit  146  operates the flow management center  82  and compressor  76 . This may be for example an interface circuit that connects to the electric motor for driving the compressor  76  in response to system temperature inputs. The interface circuit may also provide a duty cycle signal for controlling the expansion valve  116  to maintain a regulated average superheat temperature in the compressor suction line. Additionally, the driver  146  may include an interface circuit coupled to four-way switch  78  for reversing the switch from cooling mode to heating mode. 
     Processor  140  includes a main program  151 , depicted in the flowchart of FIG. 6, to control the operating mode selection, compressor speed control, and electronic expansion valve (EXV) control. FIG. 6 gives an overview of the control strategy illustrating the major functional modules that are involved. 
     Referring to FIG. 6, the main program  151  is illustrated. The main program  151  provides the timing for execution of the various control modules. At step  152  the program enters the operating mode selection module in which the operating mode of the system is selected. The supported operating modes include defrost mode, vent mode, PTC heater mode, heat pump mode, and air conditioning mode. The inputs monitored by the controller  130  to select the HVAC system  50  operating mode include the position of mode selector switch  72 , temperature lever  74 , inlet temperature, and during a heating operation the capacity of heat pump mode. Although the preferred embodiment has five discrete operating modes, the principles of the invention can be extended to systems having either fewer operating modes or a continuously variable set of operating modes. 
     FIG. 16 illustrates the system operating modes. During the PTC heater/defrost mode, when the ambient temperature is less than 40° F., controller  130  turns on the PTC heater  62  and moves the blend door to a position determined by the location of temperature lever  74 . However, for the first 3% of temperature lever  74  travel from the full cold position the controller turns off PTC heater  62  and only enables the vents. 
     In the heating mode, with ambient temperatures greater than 40° F. or defrost operation with ambient temperatures between 40° F. and 60° F., controller  130  turns on the heat pump and if necessary the PTC heater with blend door to generate the desired temperature that is reflected by the position of temperature lever  74 . For the first 3% of temperature lever  74  travel from full cold the controller  130  turns off the heat pump and PTC heater  62  and only enables the vents. At temperatures greater than 100° F. the controller  130  turns off PTC heater  62 . 
     The third operating mode, cooling mode, is selectable for ambient temperatures that are greater than 40° F. Cooling mode is also used for defrost when the ambient temperature is greater than 60° F. For the first 33% of temperature lever travel the controller  130  varies the compressor suction pressure set point from 20 to 45 psig as the temperature lever  74  is moved from cold to warm. Varying the suction pressure set point causes a direct change in the compressor speed, thereby causing the air temperature at the duct outlets to change. From 33% to 100% of temperature lever travel the controller  130  sets the compressor  76  suction pressure to a constant 30 psig and turns on the PTC heater  62  to reheat the conditioned air. 
     Retuning to FIG. 6, at step  154  the program enters the recirculation door positioning module which is described below with reference to FIG.  13 . The recirculation door positioning module controls the proportion of fresh air to recirculated air that is blown into the passenger compartment. At steps  156  and  158  the program enters modules for monitoring and disabling the compressor in response to detected faults. The compressor speed control module, which is described below with reference to FIG. 8, is entered at step  160 . Varying the speed of compressor  76  causes a proportional change in the air temperature blown from the duct outlets  64 - 68 . Step  162  leads to the EXV control module which is described with reference to FIG.  7 . The EXV control module  162  modulates the output of the expansion valve  116  in response to changes in the vapor temperature sensed at the compressor  76  and the compressor suction pressure. Each of the above listed modules will now be further explained. 
     FIG. 7 illustrates the detailed operation of EXV control module  162 . The module  162  controls the volume of refrigerant that is pressure reduced by the expansion valve  116  to maintain a relatively constant superheat temperature at the outlet of the evaporator. As low-pressure refrigerant flows from the expansion valve  116  through the evaporator it absorbs heat from the air passing through the evaporator. After absorbing sufficient heat the low-pressure refrigerant transitions to a vapor state. Any further heat that is absorbed by the vapor raises the refrigerant temperature above the saturation temperature into a superheated temperature region. To reduce the outlet temperature of the refrigerant the volume of refrigerant flowing into the evaporator is increased, thereby increasing the heat load capacity of the refrigerant. However, if there is too great a volume the refrigerant will not transition to the vapor state, resulting in the compressor  76  being swamped by liquid refrigerant. An insufficient volume of refrigerant flowing into the evaporator results in the refrigerant transitioning to the vapor state before reaching the outlet of the evaporator. Vapor state refrigerant has less capacity to store heat energy than liquid state refrigerant, therefore the portion of the evaporator that contains vapor state refrigerant has less capacity to store heat energy, reducing the efficiency of the evaporator. It is desirable to control the EXV  116  such that the liquid to vapor transition occurs slightly before the outlet of the evaporator causing the refrigerant to superheat a predetermined amount. This maximizes the efficiency of the evaporator by ensuring that virtually the entire coil is used for absorbing heat. 
     In step  164  the proportional-integral-differential (PID) constants are chosen based upon whether the system is in heating mode or cooling mode. The selection of PID constants is based upon the particular system characteristics and is well known in the art. Following selection of the PID constants the EXV control module proceeds to steps  166  and  168  wherein the expansion valve duty cycle is initialized based upon ambient temperature and operating mode when the system first enters either heat pump mode or air conditioning mode. The graph appended to step  168  depicts the selection criteria for the duty cycle. Ambient temperature is sensed by temperature probe  135  located in front of the outside heat exchanger  80 . The initial duty cycle is then set to a value ranging from 50% to 100% of the maximum EXV duty cycle depending on the ambient temperature. After setting the initial expansion valve duty cycle the system transitions through a start-up period before settling into steady-state operation. 
     During steady-state operation the duty cycle of the EXV is varied in order to maintain a constant superheat temperature, 4° F. greater than the saturation temperature, at the inlet to compressor  76 . At step  170  the average superheat temperature is calculated by measuring the vapor temperature of refrigerant exiting the evaporator and subtracting the saturation temperature of the fluid. The saturation temperature is obtained by measuring the compressor inlet suction pressure and using the saturation temperature that corresponds to the suction pressure. Although the present embodiment of the invention calculates the average superheat temperature from the vapor temperature and suction pressure, it is within the scope of the invention to use the vapor temperature with an evaporator inlet temperature including compensating for the evaporator pressure drop. The outlet of the expansion valve  116  located in the flow management center provides a common temperature measurement location for evaporator inlet temperature in either heating mode or cooling mode. In conventional systems that use the evaporator inlet temperature to calculate the superheat temperature; temperature probes are required at the inlets to both the inside and outside heat exchangers to provide inlet temperature in both operating modes. 
     The updated superheat temperature from step  170  is used at step  172  to calculate a revised setting for the EXV duty cycle. As a final step, at step  174  the controller  130  limits the value of the EXV duty cycle to between 5% and 100% to ensure the device remains within a known operating region. 
     Referring to FIG. 8, the compressor speed control module  160  is illustrated. The compressor speed is controlled by applying a variable duty cycle to the electric motor that drives the compressor  76 . The duty cycle is varied in response to a controlling input such as temperature lever position and compressor suction pressure. Varying the speed of compressor  76  causes a proportionate variation in the discharge temperature and discharge pressure of refrigerant flowing out of the compressor  76  as well as an inversely proportional change in the compressor suction pressure and refrigerant suction temperature. The increased refrigerant discharge temperature results in an increased condenser temperature, increasing the capacity of the HVAC system  50  to provide heat during heating mode. The decreased refrigerant suction temperature results in a decreased evaporator temperature, increasing the capacity of the HVAC system  50  provide cooling during cooling mode. The speed of the compressor  76  is therefore varied to maintain air blown into the passenger compartment at a relatively constant temperature during both heating mode and cooling mode. 
     The desired temperature is set by adjusting the temperature lever  74  on front panel  55 . The controller  130  calculates the target suction pressure corresponding to the temperature lever position (x/L) which is equal to 20+75*(x/L) for a lever travel distance equal to 33% of the available distance. Using the suction pressure as the controlled parameter instead of air temperature provides a more stable and faster responding system. 
     Conventional systems that use air temperature as the controlled parameter have problems with surging of the compressor  76  in addition to slow response time. As the sensed outlet air temperature changes due to transient effects including changes in vehicle speed or passing through intermittent sunlight, the compressor speed is changed in an attempt to keep the outlet temperature constant. When the compressor speed is constantly changing the passenger perceives the changes as surging in the propulsion of the vehicle. In the preferred embodiment, the EXV control loop regulates a constant outlet temperature while the compressor regulates a constant suction pressure. As the outlet air temperature changes the heat that is transferred between the refrigerant and the inside heat exchanger varies, causing the refrigerant superheat temperature to change. In response to the change in the superheat temperature the duty cycle of pressure reducing valve  116  is changed by controller  130 , causing a shift in the flow of refrigerant, resulting in a slight variation of the compressor suction pressure. The controller  130  then modifies the speed of compressor  76  to bring it in conformance with the target suction pressure. However, the required change in the speed of the compressor  76  is significantly less than the change that would be required in an HVAC system that uses compressor speed alone to compensate for changes in outlet temperature. The minor change in compressor speed is imperceptible to the passengers, leading to enhanced driving comfort. 
     In addition to eliminating surging, the response time of HVAC system  50  is reduced by using suction pressure as the controlled input. Cooling air at the desired temperature is blown over passengers in significantly less time than conventional systems that control air temperature directly. As a result, unlike conventional systems, PTC heating of the cooled air is not required to provide fine control over the air temperature, resulting in more energy efficient vehicle operation. 
     During heat mode the compressor speed is varied in reaction to changes in the temperature of the air flowing out of the inside heat exchanger  88 . In heat pump mode, unlike air conditioning mode, suction pressure is not directly related to the temperature of air flowing out of the inside heat exchanger. Therefore the air temperature sensed by temperature probe  133  is used as the controlling input for compressor speed. 
     In step  176  the controller  130  calculates the error and error derivative to be used in the PID controller for the controlled input. In air conditioning mode the controlled input is the suction pressure and in heat pump mode the controlled input is the post inside heat exchanger air temperature measured by temperature probe  133 . In step  178  the PID constants corresponding to the appropriate operating mode are selected. Then in step  180  the PID controller calculates the change in compressor duty cycle based on the PD constants and the calculated error and error derivative. The revised duty cycle is limited to between 5% and 90% to ensure the compressor  76  is operated within specified parameters. 
     FIG. 9 illustrates the interaction between the EXV control loop and the compressor speed control loop during the cooling mode start-up transition. As explained the EXV control loop regulates the volume of refrigerant that flows through pressure reducing device  116  maintaining a predetermined refrigerant superheat temperature at the outlet of the evaporator. A secondary effect of the EXV operation is that as the EXV permits an increased-volume of refrigerant to flow, the suction pressure at the inlet to compressor  76  decreases. The operation of the compressor  76  has a corresponding interaction with the EXV. When the speed of compressor  76  is changed, the resulting change in suction pressure and temperature at the inlet to compressor  76  causes a change in the saturation temperature of refrigerant that flows through the evaporator. Increased compressor  76  speed, causes a lower suction pressure, leading to a lower saturation temperature, resulting in the refrigerant temperature rising to the predetermined superheat temperature earlier in the traverse of the evaporator. The EXV loop compensates for the change in superheat temperature by permitting an increased volume of refrigerant to flow through the evaporator, thereby causing a higher suction pressure. When the HVAC system  50  first turns on, if the pressure reducing valve  116  is set to an initial duty cycle of 0%, the volume of refrigerant flowing through the evaporator will lag the compressor speed throughout the entire start-up time period, delaying the start-up, resulting in a start-up time period of approximately 2.5 minutes. 
     Assuming an ambient temperature of 40° F., the EXV is set to an initial duty cycle of 50% at step  168  (see FIG.  7 ). The compressor suction pressure is set to achieve the target suction pressure corresponding to the location of temperature lever  74 . Initially, the compressor suction pressure decreases slightly during the first seconds of operation as fluid pours through the EXV, then as the compressor spins up towards steady-state speed suction pressure begins to increase significantly. At the same time the EXV duty cycle increases until the suction pressure has increased to a point where the EXV begins to track the suction pressure. During the early stages of start-up it is not unusual for the compressor to flood until the compressor speed increases a sufficient amount to develop appropriate suction pressure. In the preferred embodiment the compressor is operated on the borderline of flooding during the start-up transition thereby contributing to a faster system response time. Also, as the EXV duty cycle begins to track the suction pressure it will overshoot its steady-state value by a slight amount. The underdamped response displayed by the EXV control loop results in a further reduction in the system response time. In combination the improvements result in air cooled to the desired temperature blowing over the faces of passengers within approximately 35 seconds of system start-up. 
     From the foregoing it will be understood that the invention provides a system for improving the steady-state response time of an automotive HVAC system. Additionally, the invention permits a reduction in the start-up time of an automotive air conditioning system. Also, the invention provides a system for controlling an HVAC system that employs a flow management device. The invention further provides a system for controlling an HVAC system incorporating a centralized flow management center. 
     Anti-Fog System for Reversible Air Conditioning and Heat Pump HVAC System for Automobiles 
     Referring to FIG. 10, a single loop reversible air conditioning and heat pump system  191  is illustrated with the corresponding temperature cycle diagrams for air conditioning mode  190  and heat pump mode  192 . As will be described, the preferred embodiment of the present invention prevents undesirable fogging by slowly increasing the speed of compressor  76  over a predetermined period of time. Generally, in reversible HVAC systems fogging may occur during the transition from cooling mode to heating mode. Prior to describing the solution provided by the presently preferred embodiment, a brief description of the refrigeration cycle and how fogging occurs in a reversible system is provided with reference to FIG.  10 . 
     The refrigeration cycle essentially uses a small amount of energy to power a compressor in order to transfer a greater amount of heat energy from one environmental region to another environmental region. It does this by using the cooling effect of evaporation to lower the temperature of the air passing through one heat exchanger (the evaporator)  88  and using the heating effect of condensing high temperature, high pressure gas to raise the temperature of the air passing through another heat exchanger (the condenser)  80 . With reference to waveform t 1  of FIG. 10, drawn from right to left, the temperature profile of refrigerant flowing from an evaporator  88 , through a compressor  76  and four-way switch  78 , and then through a condenser  80  is illustrated. Refrigerant entering the evaporator  88  is at low pressure and low temperature. The temperature being the saturation temperature of the pressure reduced refrigerant. As the refrigerant passes through the evaporator  88  heat energy from air that is blown through the evaporator  88  is absorbed by the refrigerant. The air that exits the evaporator  88  is noticeably cooled due to the transfer of heat energy from the air to the refrigerant. The cooler air no longer has the capacity to retain the same amount of moisture as the warmer air that was blown into the evaporator  88 , therefore the excess moisture condenses out of the air onto the external surface of the evaporator  88 . The vapor state refrigerant flows from the evaporator  88  to the compressor  76  where it is compressed to a high pressure, high temperature vapor before flowing into the condenser  80 . 
     When the controller  130  commands a change to heating mode the orientation of four-way switch  78  is changed, thus interchanging the refrigerant connections to the compressor  76 , thereby reversing the flow of refrigerant through the system causing the heat exchangers to change functions. Referring to waveform t 2  of FIG. 10, drawn from left to right, pressure reduced refrigerant flowing into outside heat exchanger  80  (the evaporator) absorbs heat energy from the outside air which is blown through the evaporator  80 . The refrigerant flowing through the evaporator remains at its saturation temperature for a majority of the traverse. As the refrigerant nears the end of the evaporator  80  the accumulated heat energy that has been absorbed causes the refrigerant to transition to a vapor state. Any further heat energy that is absorbed in the refrigerant causes the refrigerant temperature to increase beyond the saturation temperature into a superheated temperature range. The superheated refrigerant flows to the compressor  76  which compresses it to a high pressure, high temperature vapor which is directed to the inside heat exchanger (the condenser)  88 . As the high temperature vapor flows into the condenser  88 , the temperature of the condenser  88  rapidly rises to an equivalent temperature. Moisture that had accumulated on the inside coil  88  during the air conditioning mode begins to boil off as the condenser  88  increases in temperature. The moisture is absorbed by air flowing through condenser  88  into the passenger compartment. Fogging then occurs when the moisture laden air strikes the cold inside surface of the passenger compartment windows. 
     FIGS. 5 and 11 illustrate an exemplary anti-fogging system for controlling the operation of a reversible HVAC system  50  for automobiles. FIG. 5 as explained earlier in this specification illustrates a control system for an automotive HVAC system. Using the same hardware configuration, controller  130  minimizes the effects of fogging by gradually increasing the compressor speed at a predetermined rate and regulating the flow management center operation to ensure efficient use of the evaporator. Although a flow management center  82  is employed in the preferred embodiment it is within the scope of the invention to use a pressure reducing device with a separate receiver/drier. Additionally, the invention encompasses any variable speed or capacity compressor, even though the compressor in the preferred embodiment is an electric compressor. 
     Processor  140  is programmed to control the compressor speed and flow management center operation as depicted in the flowchart of FIG.  11 . FIG. 11 provides a general overview of the main system operating modes and the detailed program steps related to the anti-fogging routine. In the preferred embodiment of the invention the steps that are included in the anti-fogging routine  201  are spread throughout a number of program modules such as the operating mode selection  152 , compressor speed control  160 , and EXV control  162  (see FIG.  6 ). Calculated changes to the outputs that control the speed of compressor  76  and the regulation of pressure reducing device  116  only occur within the designated modules. To clarify the included steps, they have been brought together and listed in anti-fog routine  201 . 
     At step  200  the program enters air conditioning mode in which cooling air is blown into the passenger compartment. During air conditioning mode, as a byproduct of the refrigeration process moisture accumulates on the external surface of inside heat exchanger  88 . At step  202  an anti-fog flag is set to provide an indication that there is moisture on the surface of the inside heat exchanger  88 . The anti-fog flag will remain set until heat pump mode is entered at step  204 . At step  206  the program continues into the anti-fog sequence  208  if the anti-fog flag is set, otherwise it branches off to steady-state heat pump mode at step  210 . 
     The anti-fog sequence begins with selecting a post-inside heat exchanger air target temperature and a duration of operation at step  212  from a table of values that are represented in the graph. The actual post-inside heat exchanger air temperature is measured by probe  133 . The target temperature is set equal to the ambient plus an offset that is increased over time. Limiting the post-inside heat exchanger target temperature to a specified offset above ambient indirectly limits the temperature of the compressed refrigerant vapor that flows into the condenser  88 . The evaporation rate of moisture located on the inside heat exchanger  88  is directly related to the refrigerant temperature at the inlet to condenser  88 . Therefore, gradually increasing the target temperature causes a gradual increase in the compressor speed, which causes a gradual increase in the compressor discharge pressure, which results in a gradual increase in the refrigerant temperature at the inlet to the condenser, thereby limiting the evaporation rate of moisture on the condenser  88 . 
     At step  214  the compressor target suction pressure is set to 45 psi. Starting the suction pressure at 45 psi ensures that the starting discharge pressure and temperature are low enough to prevent uncontrolled moisture evaporation from the condenser  88 . The suction pressure is related directly to the speed of compressor  76 . 
     At step  216  a PID controller calculates the new compressor speed setting based upon the target temperature and previous suction pressure. The change in suction pressure from the previous setting is limited to prevent undesirable changes in compressor speed which could lead to high discharge temperatures and uncontrolled condenser moisture evaporation. Although the preferred embodiment of the invention controls the compressor speed to regulate the moisture evaporation rate, it is within the scope of the invention to control other system parameters such as suction pressure, discharge pressure, or condenser inlet temperature. 
     If the post-inside heat exchanger target temperature is less than the target temperature that correlates to the temperature lever  74  position, then the PTC heater  62  is turned on and the blend door  61  is set to a position that will enable the HVAC to achieve the temperature lever target temperature. The required door  61  position is obtained from a lookup table that correlates blend door position to differential temperature and airflow. 
     At step  218  the recirculation door  60  is set to the full fresh air position. Setting the recirculation door  60  to the full fresh air position in combination with slowly evaporating moisture from the condenser prevents fogging in the passenger compartment. As moisture is slowly evaporated off of the condenser it is absorbed by the fresh air flowing past the recirculation door  60 , through inside heat exchanger  88 , and into the passenger compartment. The moisture laden air flowing into the passenger compartment from the outside causes the internal air pressure to increase, acting to drive air out of the compartment through vents and other unsealed openings. Pushing air out the vents prevents an excessive amount of moisture laden air from accumulating in the passenger compartment as well as ensuring that the driest possible air is passed over the inside heat exchanger  88 . 
     The anti-fog sequence continues until controller  130  has executed the table of values depicted graphically at step  212 . Having completed the predetermined routine, all of the moisture that existed on inside heat exchanger  88  has evaporated and therefore the temperature of the refrigerant entering the condenser  88  no longer needs to be controlled. The anti-fog flag is reset and the heat pump system transitions to normal steady-state heat pump mode in which the speed of the compressor  76  is controlled such that a desired duct outlet temperature as selected with temperature lever  74  is attained. 
     From the foregoing it will be understood that the invention provides a system which controls fogging when changing modes in a reversible HVAC system. Additionally, through the use of the anti-fogging method the rate of initial heating of the passenger compartment is not compromised. Additionally, the invention permits a system which controls fogging in an HVAC system when initially starting air conditioning mode. 
     Heating System in a Reversible Air Conditioning and Heat Pump HVAC System for Electric Vehicles 
     FIGS. 5 and 12 illustrate an exemplary temperature control system for a reversible air conditioning and heat pump HVAC system for an electric automobile. FIG. 5 illustrates the interconnection of controller  130  to an automotive air conditioning circuit  50 . Controller  130  controls the compressor speed, flow management center  82  operation, and blend door  61  positioning based upon inputs from front panel  55 , duct  56 , and the refrigerant system. The controller  130  is preferably a microprocessor-based circuit, that includes processor  140  for executing a program, its associated memory  142 , an A/D  144  for converting analog signals into digital inputs, and a driver circuit  146  for interfacing with system components. 
     Processor  140  is programmed to control the heating mode selection that is depicted in the flowcharts of FIGS. 12A and 12B. The heating mode selection programs control the operation of the HVAC circuit  50  during a heating operation. In the preferred embodiment of the invention the steps that are included in the heating mode selection modules are spread throughout a number of program modules such as the operating mode selection  152 , compressor speed control  160 , and EXV control  162  (see FIG.  6 ). Calculated changes to the outputs that control the speed of compressor  76  and the regulation of pressure reducing device  116  only occur within the designated modules. To clarify the included steps, they have been brought together and listed in the two heating mode selection modules. 
     Heat to the passenger compartment is provided by a combination of the HVAC in heat pump mode and PTC heaters  62  depending on the ambient temperature and the requested target temperature as selected by the position of the temperature lever  74 . 
     For ambient temperatures less than 40° F. heat is supplied only by the PTC heater as the reversible HVAC refrigerant system is disabled to prevent icing of the heat exchangers  80  and  88  which would result in reduced airflow and odors in the passenger compartment. At ambient temperatures greater than or equal to 40° F. heat is supplied by either the heat pump, the PTC heater  62 , or the heat pump supplemented by the PTC heater  62 . 
     Referring to FIG. 12A, at step  270  a target temperature is calculated based upon the position of temperature lever  74 . A lookup table contains values that correlate temperature lever position to the target temperature of the air flowing from the duct outlets  64 - 68 . The creation of a lookup table containing such values is well known in the art. At step  272  the target temperature is then compared to the temperature of air flowing into inside heat exchanger  88 . The pre-indoor heat exchanger air temperature is measured by probe  132 . If the air temperature at probe  132  exceeds the target temperature the PTC heater  62  is turned off, the heat pump is turned off, and the blend door  61  is set to the max cool position. In the max cool position air bypasses the PTC heater and flows directly to the duct outlets. During this mode of operation the outside air which flows into the duct  56  is warmer than the passenger has requested via the temperature lever  74 . To cool the incoming air to the desired temperature the passenger has the option of enabling air conditioning mode. 
     For incoming air that is colder than the target temperature the compressor speed is adjusted by a PID controller at step  276  to drive the temperature of post inside heat exchanger air to the target temperature. As compressor speed is increased the refrigerant suction pressure and temperature decreases enabling the refrigerant to absorb a greater amount of heat from the external air as the refrigerant traverses the outside heat exchanger (evaporator)  80 . The refrigerant is additionally compressed by the compressor to a greater discharge temperature and pressure prior to being routed to the inside heat exchanger (condenser)  88 . The increased heat load of the refrigerant, obtained from the outside heat exchanger  80 , is then transferred to the air flowing through the inside heat exchanger  88 . The increased heat transfer causes a commensurate increase in the post inside heat exchanger air temperature, assuming the ambient temperature and air flow rate remains constant. 
     At step  278  the post inside heat exchanger air temperature is measured by probe  133  and compared to the target temperature. The post inside heat exchanger air temperature represents the air temperature prior to the PTC heater. If the air temperature is greater than the target temperature, then supplemental heat is not required to achieve the target temperature. Therefore, at step  280  the controller turns PTC heater  62  off, sets the blend door  61  to the max cool position, and returns to step  270  to begin another iteration. This is the normal operating loop during heat mode operation as the controller  130  regulates the air temperature to the selected target temperature. The post inside heat exchanger air temperature will exhibit normal closed loop operation by fluctuating slightly about the target temperature. 
     If the measured post inside heat exchanger air temperature is less than the target temperature, then the electric heater, PTC heater  62 , is turned on. As the air flow rate across the PTC heater  62  increases, the heat output of the device increases thereby transferring a greater amount of heat to the passenger compartment. To regulate the quantity of heat that is transferred to the passenger compartment blend door  61  provides a path for a portion of the air to bypass the PTC heater  62  and recombine downstream with air that has flowed through the PTC heater  62 . By reducing the quantity of air that flows over the PTC heater  62 , less heat is transferred to the air, thereby reducing the commensurate increase in the temperature of the air, and providing a simple means of regulating the temperature of the recombined air. 
     At step  282  the required blend door position to achieve the target temperature is calculated in a manner known in the art. The required effectiveness represents the amount of PTC heating that is required to raise the temperature of the post inside heat exchanger air to the target temperature at the existing airflow across the PTC. At step  284  the controller  130  sets the position of blend door  61  and the loop returns to step  270  to start another iteration. This is the normal operating loop when supplemental heat from the PTC heater  62  is required to raise the duct outlet air to the requested temperature. Each time through steps  270 ,  272 ,  276 ,  278 ,  282 , and  284  the position of the blend door  61  is varied slightly as the controller  130  responds to changing conditions. 
     Alternatively, the heating mode selection program can be implemented as illustrated in FIG.  12 B. The program illustrated in FIG. 12B is particularly suitable for operating modes where the overhead energy that is expended turning on the heat pump or PTC heater  62  exceeds the energy required to raise the passenger compartment temperature to the desired temperature. At step  300  a forty second timer is started. The timer sets the time period during which the heat pump attempts to attain the target temperature. At step  302  the heat pump target temperature is calculated based on the position of temperature lever  74 . The compressor speed PID controller is adjusted at step  304  to drive the compressor speed towards attaining the target temperature. At step  306  the heat pump gain is calculated. The heat pump gain represents the work the heat pump contributes to raise the temperature of the passenger compartment under the existing operating conditions. The heat pump gain is set equal to the outlet temperature, probe  133 , minus the inlet temperature, probe  132 , divided by the outlet temperature. At step  308  the post-inside heat exchanger air temperature as measured by probe  133  is compared to the target temperature calculated at step  302  to determine if the heat pump is capable of attaining the target temperature. If the heat pump does attain the target temperature the forty second timer is reset at step  310  and the program returns to step  302 . Additionally, if the heat pump has not attained the target temperature but the 40 second timer has not timed out, the program returns to step  302  to continue to attempt to attain the target temperature. However, if the heat pump does not attain the target temperature within 40 seconds then at step  314  the measured values for heat pump gain and ambient temperature are stored for later use. Although, in the preferred embodiment the heat pump is allowed 40 seconds to attain the target temperature, it is within the scope of the invention that the allowed time may range from about 0 seconds to beyond 40 seconds. For example, the heat pump heating capability may be characterized by factory test or simulation and a number representative of the capability may be stored in memory for later recall to determine if the heat pump is capable of attaining a target temperature. 
     At step  316  the heating mode transitions from the heat pump to PTC heater  62  by gradually decreasing the heat pump output and increasing the PTC heater  62  over a 40 second period. Making a gradual transition enhances passenger comfort by reducing the noticeability of the change in system operation. At step  318  the stored value for heat pump gain is adjusted for changes in ambient temperature. At step  320  the revised value for heat pump gain is compared to the system gain that represents the amount of work required to heat the passenger compartment to the target temperature. If the system gain exceeds the heat pump gain, there is insufficient capacity in heat pump mode for heating the passenger compartment, therefore the program remains in PTC heat mode and returns to step  318 . If the heat pump gain exceeds the system gain, the heat pump is capable of supplying the required heat necessary to attain the target temperature. The program advances to step  322  and transitions from the PTC heater  62  to heat pump over a 40 second time period, finally returning to heat pump mode at step  300 . 
     From the foregoing it will be understood that the invention provides a system which minimizes energy consumption during a heating operation of an automotive HVAC system. Additionally, the method can be employed to dynamically update the heating mode selection as operating conditions change. Also, through the use of the method the energy efficiency of an electric vehicle is increased. Additionally, the invention provides an energy efficient method for controlling the passenger compartment temperature of an electric vehicle. 
     Air Handling for HVAC System for Electric Vehicles 
     Referring to FIGS. 5 and 13, an air handling system for an electric vehicle HVAC system is illustrated. FIG. 5 illustrates the interconnection of controller  130  to an automotive air conditioning circuit  50 . Controller  130  controls the compressor speed, flow management center  82  operation, and recirculation door  60  positioning based upon inputs from front panel  55 , duct  56 , and the refrigerant system. Recirculation door  60  may be set to any value from full fresh air, through part fresh air with part recirculated air, to full recirculated air. 
     The recirculation door control program  251  is illustrated in FIG.  13 . Although FIG. 13 depicts all of the recirculation door program components existing in a single separate program module, it is within the scope of the invention for the different elements to be spread throughout the system program. In the preferred embodiment of the invention the steps that are included in the heating mode selection module are spread throughout a number of program modules such as the operating mode selection  152  and recirc. door positioning  154  modules (see FIG.  6 ). To clarify the included steps, they have been brought together and listed in the recirculation door control module. 
     When the system is turned-on, step  250  is executed and the recirculation door  60  is set to the recirculation position. By starting in the recirculation position less energy is consumed controlling the temperature of the passenger compartment In recirculation mode, air from within the passenger compartment is routed through the inside heat exchanger  88  before being directed back into the passenger compartment. Therefore to raise the duct outlet air to the desired temperature the heat transferred from inside heat exchanger  88  only has to supplement the difference between the desired temperature and the temperature of the passenger compartment. In fresh air mode, to raise the duct outlet air to the desired temperature the heat transferred from inside heat exchanger  88  supplements the difference between the desired temperature and the temperature of the external air which is flowing into the passenger compartment. 
     Having set the recirculation door  60  to its initial position the program continues on to step  252  in which the inputs from the front panel  55  are interrogated to determine if a particular positioning of the recirculation door has been requested. If a recirculation door position change has been requested, then at step  254  the recirculation door is set to the requested position at step  254 . 
     In step  256  the program optionally begins an anti-fog sequence. As is explained above, fogging of the passenger compartment windows may occur when the reversible HVAC system  50  switches from cooling mode to heating mode. During the cooling mode cycle moisture accumulates on the external surface of the inside heat exchanger  88  which functions as an evaporator. When the HVAC switches from cooling mode to heating mode the refrigerant flowing into the inside heat exchanger  88 , which functions as a condenser, rapidly increases in temperature. As the refrigerant begins to raise the temperature of the condenser  88 , moisture that had accumulated on the inside heat exchanger  88  during the cooling mode begins to boil off. The evaporating moisture is absorbed by air flowing through condenser  88  into the passenger compartment. Fogging then occurs when the moisture laden air strikes the colder windows of the passenger compartment. 
     At step  258  the air handling procedure during an anti-fog sequence is performed. The front panel selection for the recirculation door  60  position is overridden as the door  60  is set to the full fresh air position. With fresh air flowing into the passenger compartment the air pressure within the compartment increases, forcing air out of vents and door seal cracks. As new fresh air carrying its load of moisture is blown into the passenger compartment, pre-existing moisture laden air is forced out through the vents to the outside environment. The recirculation door  60  remains in the fresh air position until the anti-fog sequence is completed, at which time the recirculation door is reset to its former position. 
     In step  260  the program begins an air blow-by sequence. When the vehicle speed exceeds a predetermined value, such as approximately 42 mph, the pressure from air flowing into the fresh air duct  59  flows not only through the blower  58 , but also back up through the recirculation air duct  57 . The air flowing back into the recirculation air duct  57  bypasses the inside heat exchanger  88  and PTC heater  62  which are downstream from the recirculation door  60 . Therefore, the air flowing back into the recirculation duct is unconditioned external air. The external air could vary from extremely cold dry air during winter months to very hot humid air during the summer months. The external air flows out of the duct inlets and directly onto the passengers in the passenger compartment. 
     At step  262  the program sets the recirculation door  60  to prevent an undesirable air blow-by event from occurring. The previous setting of the recirculation door  60  is overridden and the door is set to the full fresh air setting. The recirculation air duct  57  is blocked when the recirculation door  60  is in the full fresh air position, therefore the fresh air is forced through blower  58 , inside heat exchanger  88 , and PTC heater  62 . The fresh air is properly conditioned to the desired temperature before being blown into the passenger compartment and no air flows back through the recirculation duct  57 . Although in the preferred embodiment the recirculation door is set to the full fresh air setting it could alternately be set to the full recirculation air setting, in which case the fresh air duct  59  is blocked, preventing fresh air from flowing into the duct  56 . Additionally, although in the preferred embodiment the setting of the recirculation door  60  is independent of the prior position of the recirculation door  60 , the selection of the full fresh air setting versus the full recirculation setting could be based on the position of the recirculation door  60  prior to entering the air blow-by sequence. 
     From the foregoing it will be understood that the invention provides a system for selectively overriding the passenger air mixture selection under predetermined vehicle operating conditions to permit HVAC operating modes that enhance passenger comfort. Additionally, the system can be employed to improve vehicle performance by automatically adjusting the air mix during predetermined vehicle operating modes 
     System for Cooling Electric Vehicle Batteries 
     Referring to FIG. 14, a schematic of an automotive HVAC circuit  220  for an electric vehicle is illustrated. The HVAC circuit  220  is an alternative embodiment of the invention wherein heat from the battery pack  224  is used to supplement heating of the passenger compartment. The circuit  220  is similar to the HVAC circuit  50  illustrated in FIG. 1 with the addition of a heat exchanger circuit  222  for cooling a battery pack  224 . A heat exchanger circuit  222  communicates with auxiliary heat exchanger  92  to cool battery pack  224  and controller  130 , and includes a heat exchanger  228 , a battery pack  224 , a reservoir  230 , and a pump  232 . 
     Liquid high pressure refrigerant from flow management center  82  flows through expansion valve  226  and shut-off valve  86  into heat exchanger  228 . Although high pressure refrigerant in the preferred embodiment is obtained from flow management center  82 , it is within the scope of the invention to obtain high pressure refrigerant from other means such as a valve, a receiver/drier, or a reservoir. Additionally, although a thermal expansion valve is employed in the preferred embodiment, the principles of the invention may be readily extended to other pressure reducing means such as an electronic expansion valve. Shutoff valve  86  is included merely to show a possible method of controlling battery cooling by preventing the flow of refrigerant into heat exchanger  228 . The refrigerant outlet of heat exchanger  228  is connected to the compressor  76  suction line such that the vapor is combined with refrigerant vapor from other system evaporators prior to flowing into the compressor  76  inlet. 
     The coolant outlet of heat exchanger  228  connects to battery pack  224  which includes the vehicle energy storage batteries. Heat is generated in the batteries during energy storage and discharge cycles due to energy losses from converting chemical energy to electrical energy. Heat from the batteries is transferred through the battery pack into the coolant. The outlet of battery pack  224  connects to reservoir  230  which connects to the inlet of pump  232 . The pump  232  propels the coolant through heat exchanger circuit  222 . Coolant from the pump  232  flows through controller  130 , cooling the system electronics. The heat generated by the controller  130  is additionally transferred into the coolant. The controller  130  controls the operation of HVAC system  220 . The temperature of battery pack  224  is sensed by temperature probe  225  which provides an input to the Battery Energy Management System (BEMS)  234 . The BEMS  234  controls the operation of shutoff valve  86  in response to the temperature sensed by probe  225 . 
     FIG. 15 illustrates the operation of HVAC circuit  220 . In this embodiment controller  130  sets four-way valve  78  such that the system heating mode is operational. High pressure, high temperature refrigerant flows from compressor  76  outlet through four-way valve  78  into inside heat exchanger  88  which functions as a condenser. Liquid refrigerant flows from the outlet of condenser  88  through shut-off valve  84  into a bi-directional port of flow management center  82 . The refrigerant then splits with a portion flowing from an outlet of flow management center  82  to expansion valve  226 , and the remainder of the refrigerant flowing out of the expansion valve  116  of the flow management center  226  to outside heat exchanger  80 . Pressure reduced refrigerant flows through the outside heat exchanger  80  which functions as an evaporator absorbing heat energy from the outside air flowing through it. 
     In operation, the refrigerant that flowed from the outlet of flow management center  82  flows through expansion valve  226  and shut-off valve  86  before entering heat exchanger  228 . The pressure reduced refrigerant that flows through heat exchanger  228  absorbs heat energy from coolant that is routed through heat exchanger circuit  222 . The process by which heat energy is transferred from the coolant to the refrigerant in heat exchanger  222  is the same as what occurs in inside heat exchanger  88  the functioning of an evaporator described earlier. The coolant in circuit  222  flows through battery pack  224  absorbing heat from the vehicle batteries. The coolant then flows through reservoir  230  and pump  232  before absorbing additional heat from controller  130  prior to returning to heat exchanger  228 . Hot coolant enters the heat exchanger  228  inlet and transfers its heat energy to the pressure reduced refrigerant flowing through the refrigerant line within the heat exchanger  228 . The pressure reduced refrigerant transitions to the vapor state as it absorbs heat energy from the coolant. The vapor state refrigerant then flows through the four-way switch  78  before combining with vapor state refrigerant from outside heat exchanger  80  prior to the inlet to compressor  76 . 
     Coolant continues to circulate through circuit  222  so long as the temperature of the battery pack  224  remains above 40° F. When the battery pack  224  temperature decreases below 40° F. the BEMS  234  disables shut-off valve  86  interrupting the flow of refrigerant to the heat exchanger  228 . Coolant continues to flow through heat exchanger circuit  222  as the temperature of the battery begins to slowly increase. Once the temperature of the battery pack  224  once again rises above 40° F. the BEMS  234  enables shut-off valve  86 , reestablishing the flow of refrigerant to the heat exchanger  228  and the transfer of heat from the heat exchanger circuit  224  to the HVAC circuit  220  resumes. 
     Waveform h 1  of FIG. 15 illustrates the heat cycle of HVAC circuit  220 . Refrigerant flowing into evaporator  80  initially carries a heat load depicted as plateau  238 . As the refrigerant flows through evaporator  80  it absorbs heat energy from outside air that is blown through the evaporator  80 . Meanwhile, refrigerant flowing through heat exchanger  228  also carries a heat load depicted as plateau  238 . The refrigerant flowing through heat exchanger  228  absorbs heat energy that is transferred from the battery pack  224  of heat exchanger circuit  222 . The heat load of the refrigerant increases to plateau  242  when the vapor state refrigerant from heat exchangers  80  and  228  combines prior to compressor  76 . The refrigerant heat load further increases to plateau  244  when compressor  76  compresses the vapor state refrigerant to a high pressure, high temperature vapor. The stored refrigerant heat energy decreases to plateau  238  as the refrigerant traverses the inside heat exchanger  88  and the heat energy is transferred to air that is blown through into the passenger compartment. 
     Using waste heat from the battery pack to supplement heat energy absorbed from the outside air for heating the passenger compartment provides a number of advantages. It expands the operating conditions under which heat mode operation of the HVAC is possible by increasing the stored energy in the refrigerant. It improves the efficiency of the overall vehicle system by reducing the need to rely on electric energy to heat the passenger compartment. Where conventional systems would exhaust the battery pack waste heat to the external environment and use electric energy from the batteries to provide supplemental heat to the passenger compartment, the invention reduces the need for electrical heating by using the waste heat from the batteries to supplement the heat pump system. 
     During cooling mode the flow of the refrigerant through the main loop is reversed from heat pump mode. Heat from air passing through the inside heat exchanger (evaporator)  88  is absorbed by the refrigerant. The refrigerant flowing through local-zone heat exchanger  228  continues to absorb heat from the heat exchanger circuit  222  (refer to FIG.  14 ). The refrigerant from the local-zone heat exchanger  228  combines with refrigerant from inside heat exchanger  88  prior to compressor  76 . The refrigerant is compressed further adding to the heat load and directed to the outside heat exchanger  80  (condenser). As the refrigerant traverses the condenser  80  the combined heat load is shed to the outside air that flows through the condenser  80 . The refrigerant then flows to the flow management center  82  and then through the remainder of the circuit. 
     From the foregoing it will be understood that the invention provides a system for increasing the operating range of an automotive heat pump system. Additionally, the invention provides a system for improving the energy efficiency of an electric automobile. Further, the invention provides a system for efficiently distributing the heat energy of an electric automobile. Also, a method is presented for cooling the battery pack of an electric vehicle. 
     Advantages of the invention 
     From the foregoing it will be understood that the invention provides a flow management device with bi-directional ports in which refrigerant flowing into either port passes through an expansion valve and exits the other port. Additionally, the invention can integrate the receiver/drier function into a flow management device with bi-directional ports to provide the capability of tapping off refrigerant flow for secondary cooling circuits. Also, the present invention decreases the complexity of automotive HVAC systems by integrating a flow management device into the system to reduce the number of valves required to implement a reversible heating and cooling HVAC system. A further capability of the invention is to provide a centralized flow management center with taps for refrigerant to reduce the complexity of automotive HVAC systems that implement multi-zone control. 
     The invention provides a system for improving the steady-state response time of an automotive HVAC system. Additionally, the invention permits a reduction in the start-up time of an automotive air conditioning system. Also, the invention provides a system for controlling an HVAC system that employs a flow management device. The invention further provides a system for controlling an HVAC system incorporating a centralized flow management center. 
     The invention provides a system which controls fogging when changing modes in a reversible HVAC system. Additionally, through the use of the anti-fogging method the rate of initial heating of the passenger compartment is not compromised. Additionally, the invention permits a system which controls fogging in an HVAC system when initially starting air conditioning mode. 
     The invention provides a system which minimizes energy consumption during a heating operation of an automotive HVAC system. Additionally, the method can be employed to dynamically update the heating mode selection as operating conditions change. Also, through the use of the method the energy efficiency of an electric vehicle is increased. Additionally, the invention provides an energy efficient method for controlling the passenger compartment temperature of an electric vehicle. 
     The invention provides a system for selectively overriding the passenger air mixture selection under predetermined vehicle operating conditions to permit HVAC operating modes that enhance passenger comfort. Additionally, the system can be employed to improve vehicle performance by automatically adjusting the air mix during predetermined vehicle operating modes. 
     The invention provides a system for increasing the operating range of an automotive heat pump system. Additionally, the invention provides a system for improving the energy efficiency of an electric automobile. Further, the invention provides a system for efficiently distributing the heat energy of an electric automobile. Also, a method is presented for cooling the battery pack of an electric vehicle. 
     Although certain preferred embodiments of the invention have been herein described in order to afford an enlightened understanding of the invention, and to describe its principles, it should be understood that the present invention is susceptible to modification, variation, innovation and alteration without departing or deviating from the scope, fair meaning, and basic principles of the subjoined claims.