HEATING AND COOLING SYSTEM FOR AIR SPACE IN A BUILDING

Abstract: A reversible refrigeration plant serves as an air conditioner for cooling air or alternatively as a heat pump for heating air. Functions of a refrigerant evaporator and condenser are reversed in the different modes. A main stream of inlet air is divided into separate main and bypass streams of air. The main stream is heated by the condenser of the heat pump and flow between the main stream and the bypass stream is varied in an inverse ratio so that as flow through the bypass duct increases, flow through the main duct decreases and vice versa. Also, as a desired air space temperature is approached, preferably heat rejected at the condenser is gradually reduced by reducing refrigerant flow therethrough by modulating compressor suction. Also, energy is saved in both the heating and cooling modes by locating a heat exchange coil within an exhaust flow of air discharged from the space to be heated or cooled.

Description: HEATING AND COOLING SYSTEM FOR AIR SPACE IN A BUILDING BACKGROUND OF THE INVENTION The invention relates to a heating and cooling system for air, for example for maintaining a desired air temperature in a building using fresh air or recirculating room air, by using an air-conditioning system which can be "reversed" to function as a heat pump. It is known that a conventional refrigeration system using a vapour compression refrigerator can be reversed to function as a heat pump. A conventional refrigerator cools a space by extracting heat therefrom by evaporation of a refrigerant or working fluid, and rejects the heat outside by condensation. On the other hand, a heat pump utilizes the heat rejected by condensation of the refrigerant to heat a space, and extracts heat from outside by evaporation of the refrigerant. A refrigeration or heat pump plant is said to be in balance when rates of evaporation and condensation are equal. This requires matching the capacity of the evaporator and condenser as closely as possible, and this matching can present problems. In air-conditioning systems, most of the matching problems arise from the wide variations that can be experienced by the condenser when it is located in ambient air outside the building. Many methods of bringing into balance the rates of evaporation and condensation are available for air-conditioners as will be described. On the other hand, while heat pumps are well known for their relatively high cycle efficiency, problems have arisen in matching the heat rejection capacity of the condenser with the heat absorbing capacity of the evaporator. If the range of outside air temperature passing through the condenser is relatively narrow, matching of the evaporator and condenser capacities is relatively easy and the heat pump can function efficiently. However, if there is a relatively wide range of temperature differences across the condenser of the heat pump, matching of the condenser and evaporator capacities while maintaining high efficiency and high discharge temperature becomes very difficult. Matching of the capacities of the evaporator and condenser in a refrigerating unit has been accomplished by many different approaches. For a proper balance, refrigerant pressure within the system must be controlled accurately, and this can be done by controlling the rates of condensation and evaporation, using many different methods. One means of approaching balance for common, commercially produced air-conditioners when used in a temperate climate where less condenser capacity is required is to partially flood the condenser with liquid refrigerant, thus effectively reducing condenser area to compensate for cooler ambient air. On the other hand, when attempting to match capacities of the condenser and evaporator in a heat pump system when temperature of air entering the condenser varies, as in fresh air heating, many difficulties can arise. If the heat pump system has only a single compressor, unloaders using a bypass regulator are required to short circuit refrigerant, thus reducing heat capacity demands but this approach reduces efficiency when attempting to maintain adequate pressure in the condenser. Constant volume air-conditioners using make-up air have to accommodate wide variations of ambient temperature and still provide a desired cooled space temperature. Clearly, as the ambient temperature rises, more heat must be removed from the air entering the evaporator than with lower ambient temperatures. In order to accommodate these outside air temperature variations in an air conditioner, air flow control means are provided to regulate flow of outside air through the condenser coil of the air-conditioner. The higher the ambient temperature, the more air must pass through the condenser coil and vice versa. Usually air flow control means, termed face dampers, are fitted to control air flow through the main condenser coil. For relatively cool ambient temperatures, where not much heat is removed from incoming air, the face dampers will be closed so that little of the outside air will pass through the condenser. As ambient or outside temperature rises, the face dampers will be opened, to ensure that more outside air will be passed through the condenser. In all air-conditioning systems known to the inventor, the condenser used to reject heat is located in ambient air, and, to improve heat transfer, the condenser can be cooled by many different methods. The efficiency of heat transfer from the hot refrigerant in the condenser to ambient air is dependent, among other things, on temperature difference between the hot refrigerant and the ambient air. Various means, such as sprayed water cooling for direct evaporation on the coil have been used to assist in cooling the refrigerant. When an air-conditioning system is used as a heat pump, additional difficulties have been encountered in matching the capacity of the evaporator and condenser, and rate of refrigerant flow through the heat pump. Typical problems encountered include attempting to maintain design condensing pressure for all outside temperatures with a constant volume make-up air passing through the condenser coil. SUMMARY OF THE INVENTION The invention reduces the difficulties and disadvantages of the prior art by providing a relatively low capital cost heating and cool .ιg system which also operates efficiently in both heating and cooling modes. The invention also provides a relatively simple means of reducing differences in heat transfer capacity between the evaporator and condenser of a heat pump by controlling essentially constant volume fresh air flow therethrough automatically to maintain high efficiency for a relatively wide range of outside temperature. Also, the efficiency of heat transfer relative to an exhaust coil is improved by locating the exhaust coil actually in exhaust air flow from a building being cooled or heated, as opposed to locating the coil in ambient air which requires an auxiliary fan or other means. In air-conditioning, because air exhausted from the building being cooled is usually cooler than ambient, an improved temperature range for the condenser is attained which, by itself, improves heat rejection in the condenser. On the other hand, in the heat pump mode, because air exhausted from the building being heated is usually warmer than winter outside air, locating the evaporator in exhaust air flow from the building being heated will recover more heat in the evaporator than the prior art location of the evaporator in outside air. This additional heat extracted from the exhaust air can then be "intensified" by the compressor before being rejected into the fresh air stream by the condenser coil. A method according to the invention is for heating air and comprises the steps: generating main and bypass streams of air, and initially maintaining the two streams of air separate from each other, directing the main stream of air through a condenser of a vapour compression refrigerant plant operating as a heat pump to heat the main stream of air, generating a first control signal representing a difference between an actual first reference temperature and a pre-determined first reference temperature, controlling flow of the bypass stream of air in response to the first control signal so that if the first control signal reflects an actual first reference temperature below the pre-determined first reference temperature, bypass flow is increased, and if the first signal reflects an actual first reference temperature above the pre-determined first reference temperature, the bypass flow is decreased, mixing the main stream of air after passing through the condenser with the bypass stream of air, if any air were passed through the bypass flow control means, prior to delivering the mixture of air streams to the space to be heated. Preferably, the main stream of air is controlled also in response to the first control signal so that as flow through the bypass duct increases, flow through the main duct decreases and vice versa. The method also includes generating a second control signal reflecting difference between an actual second reference temperature and a pre-determined second reference temperature and gradually reducing heat rejection at the condenser as the actual second reference temperature approaches the pre-determined second reference temperature. Preferably, the heat rejection at the condenser is reduced by reducing heat gathered in an evaporator of the heat pump from air flowing through the evaporator. This diminishes the refrigeration effect, and can be accomplished by raising the boiling temperature of refrigerant entering the evaporator by raising the pressure thereof. The resultant warmer evaporator gathers less heat as the temperature difference between the evaporator and the air flowing through the evaporator becomes less. An apparatus according to the invention is for heating air and comprises a supply chamber, propulsion means, a condenser, a first temperature sensor means, a bypass flow control means, an actuator means and mixing means. The chamber has main and bypass ducts for receiving and initially maintaining separate main and bypass streams of air. The propulsion means is for generating air flow through at least one of the ducts. The condenser is of a vapour compression refrigeration plant provided in the main duct to heat the main stream of air passing through the main duct. The first temperature sensor means generates a first control signal representing difference between an actual first reference temperature and a predetermined first reference temperature. The bypass flow control means cooperates with the bypass duct to control flow of the bypass stream of air therethrough. The actuator means is responsive to the first control signal and is operatively coupled to the bypass flow control means to actuate the control means so as to vary an amount of restriction to air flow through the bypass duct, the actuator means being responsive to the first control signal. The mixing means is for mixing together the streams of air discharged from the main and bypass ducts prior to discharging the mixed streams of air into the space to be heated. Preferably, the apparatus further comprises a main flow control means cooperating with the main duct to control flow therethrough and coupling means coupled to the main flow control means and the bypass flow control means so that the main flow control means is actuated in an inverse relationship to the bypass flow control means. The apparatus further comprises a second temperature sensor means and a heat rejection reducing means. The sensor means generates a second control signal representing difference between an actual second reference temperature and a pre-determined second reference temperature. The heat rejection reducing means gradually reduces heat rejected at the condenser by cooperating with the condenser so as that as the actual second reference temperature approaches the pre-determined second reference temperature, heat rejected to air passing through the condenser is gradually reduced by reducing flow of refrigerant through the condenser. Another method of heating air according to the invention comprises the steps of: directing a stream of air through a condenser of a vapour compression heat pump to heat the stream of air, generating a control signal representing difference between an actual reference temperature and a pre-determined reference temperature, gradually reducing heat rejection at the condenser as the actual reference temperature approaches a pre-determined reference temperature. Preferably the heat rejection at the condenser is reduced by reducing heat gathered in the evaporator of the heat pump from air passing through the evaporator. Another apparatus for heating air according to the invention, comprises a supply chamber, a condenser, a temperature sensor and heat rejection reducing means. The supply chamber has a duct to receive air to be heated and a propulsion means for generating flow of a stream of air through the duct. The condenser is of a vapour compression heat pump and is provided in the duct to heat the stream of air passing along the duct. The temperature sensor means is for generating a control signal representing difference between an actual reference temperature and a pre-determined reference temperature. The heat rejection reducing means is to gradually reduce heat rejected at the condenser and cooperates with the condenser so that as the actual reference temperature approaches the pre-determined reference temperature, heat rejected to air passing through the condenser is gradually reduced. Preferably, the heat rejection reducing means is a fluid flow modulating valve cooperating with a conduit communicating with an intake of a compressor of the heat pump. The fluid flow modulating valve is responsive to the temperature sensor so as the actual reference temperature approaches the pre-determined reference temperature, flow of refrigerant through the condenser is reduced so that heat rejected at the condenser is reduced. Yet another method of heating an air space according to the invention comprises the steps of: drawing air through a condenser of a heat pump so that condensation of refrigerant therein heats the air prior to discharging the air into the air space, drawing air from the space to be heated through an evaporator of the heat pump to heat the refrigerant therein, transferring the heated refrigerant from the evaporator to the condenser. Yet another apparatus according to the invention is for heating an air space and comprises an inlet duct and an outlet duct communicating with the space, and means to move the air through the inlet duct into the space, and to discharge air from the space through the outlet duct. The apparatus also includes a refrigerant condenser in the inlet duct, a refrigerant evaporator in the outlet duct, and refrigerant conduit means and a compressor communicating the evaporator and the condenser and transmitting refrigerant therebetween to function as a heat pump. Yet a further method according to the invention is for cooling an air space and comprises the steps of: drawing air through an evaporator of a refrigeration plant so that evaporation of the refrigerant thereof cools the air prior to discharging the air into the space, drawing air from the space to be cooled through a condenser of the refrigeration plant to cool the refrigerant thereof, transferring the cooled refrigerant from the condenser to the evaporator. Yet a further apparatus, according to the invention, is for cooling an air space and comprises an inlet duct and an outlet duct communicating with the space, and means to move the air through the inlet duct into the space and to discharge air from the space through the outlet duct. The apparatus further includes a refrigerant evaporator in the inlet duct, a refrigerant condenser in the outlet duct and refrigerant conduit means and a compressor communicating the evaporator and condenser and transmitting refrigerant therebetween to function as a refrigeration plant. A detailed disclosure following, related to drawings, describes several preferred apparatus and methods according to the invention which are capable of expression in structure and methods other than those particularly described and illustrated. DESCRIPTION OF THE DRAWINGS Figure 1 is a simplified fluid schematic showing a refrigerant circuit of a heating and cooling system according to the invention, also showing some peripheral control circuits, temperature sensors and schematic connections. Figure 2 is a simplified diagram of an air flow circuit of the apparatus according to the invention, the apparatus being shown in a one hundred percent fresh-air heat pump mode, Figure 3 is a simplified diagram of the air flow circuit of the apparatus according to the invention, the apparatus being shown in a one hundred percent re-circulating heat pump mode, Figure 4 is a simplified top plan as seen from lines 4-4 of Figures 5 and 6, shown partially in section, of a coupled system embodiment according to the invention showing relative location of various components, many other components being omitted for clarity, Figure 5 is a simplified elevation and section of the apparatus as seen from line 5-5 of Figure 4 , some portions being omitted for clarity, the invention being shown in the one hundred percent re-circulating heat pump mode. Figure 6 is a simplified elevation of the apparatus as seen from lines 6-6 of Figures 4 and 5, some portions being omitted for clarity. Figure 7 is a simplified longitudinal section through a coupling device according to the apparatus, the device being shown in an extended or uncoupled configuration. Figure 8 is a simplified section on line 8-8 of Figure 7, Figures 9, 10 and 11 are simplified diagrams showing actuation of the coupling device and dampers. Figure 12 is a simplified electrical schematic showing a portion of electrical control means, some significant signals and some connections to other components. DETAILED DISCLOSURE Figure l A heating and cooling apparatus 10 according to the invention resembles an air-conditioning system for cooling air, and can also be reversed to function as a heat pump to heat the air. The terms "heat pump", "refrigeration plant" or "air-conditioner" as used herein are used synonymously to describe a vapour compression machine which can be operated in either mode, to deliver heat to or to extract heat from air. The apparatus 10 has a motor/compressor assembly 12 complete with an internal line-brake overload device, crankcase heater and other features which are not shown but are common to refrigeration plant compressors. The apparatus includes generally similar heat exchange coils 15 and 16, the coil 15 being for exhaust air which is returned to outside, and the coil 16 being for supply air drawn into the building from outside. Each oil is relatively conventional, and is fabricated from copper or aluminum tubes with fins, with a distributor on a counter flow side and a header on an opposite side. Each heat exchange coil is fitted within a refrigerant circuit with components as follows, most of which are also found in conventional refrigeration plants or air-conditioning systems. Direction of flow in the refrigerant circuit of Figure 1 is designated by three different types of arrows. A large or multi-bodied arrow shows direction of flow which is independent of the mode c" operation, in other words the direction of flow in conduits designated by a multi-bodied arrow does not change as the mode of operation changes. A second type of arrow is a single-bodied full outline arrow which shows direction of flow in a heating mode only. A third type of arrow is shown in broken outline, shows a reversed direction of flow which occurs only in a cooling mode. As refrigerant flow direction in the cooling mode follows conventional air-conditioning practice, it is not described in detail. Instead, flow in the heating mode is described, and thus refers to the arrows in full outline, namely the multi-bodied arrows and the single-bodied full outline arrows. In the heating mode, the exhaust coil 15 acts as an evaporator, and thus receives liquid at an inlet thereof, and exhaust vapour at an outlet thereof. Conversely, the supply coil 16 acts as a condenser and receives vapour at the inlet thereof and exhausts liquid at the outlet. A first liquid line 19 feeds liquid into the exhaust air coil 15 and a vapour line 21 leaves the coil. An exhaust coil thermostatic expansion valve 23 is fitted in the line 19 and an exhaust coil check valve 25 is coupled in parallel with the valve 23. The valve 23 is connected by a capillary tube 24 to a temperature sensing bulb 26 on the line 21 at the outlet of the coil 15, the valve 23 thus being responsive to the coil outlet temperature. A second liquid line 27 extends from a junction 24 with the first liquid line 19 and feeds liquid to a de-superheating thermostatic expansion valve 28, and then feeds liquid into a pressure vessel 30 with an internal J-tube 32 on an outlet connection. The vapour line 21 feeds vapour into a hermetic, pilot-operated, slide-type, four-way reversing valve 34 which communicates with three other lines, namely a refrigerant suction line 36, a refrigerant vapour line 38 and a discharge line 40 extending from the outlet of the compressor. In the heating mode as shown, the valve 34 is positioned so that the vapour line 21 communicates with the refrigerant suction line 36, and the discharge line 40 communicates with the vapour line 38. The line 36 communicates with a modulating suction cut-off valve 37 which is electronically controlled by means to be described, and, when required, can throttle refrigerant as it passes along the line 36. One example of the valve is manufactured by the Sporlan Valve Company of St. Louis, Missouri, U.S.A., and is called a CDA electronic temperature control valve and is a solenoid-actuated, flow modulating valve. This is an important valve for the invention and, when required, can throttle supply of the vapour to the compressor, as will be described, from full flow to about 40% of normal flow. The J-tube 32 in the pressure vessel 30 communicates with a suction line 42 which feeds vapour into the compressor 12 and communicates with the discharge line 40. A temperature sensing bulb 44 on the line 40 controls the thermostatic expansion valve 28 through a capillary tube 45. The valve 28 limits compressor discharge temperature by metering liquid refrigerant directly into the final compressor suction line 42. The refrigerant vapour line 38 feeds vapour into the coil 16 for heating the supply air by condensation. The coil 16 communicates with a thermostatic expansion valve 46 mounted in parallel with a supply coil check valve 48. A temperature sensing bulb 47 in the line 38 controls the valve 46 through a capillary tube 49. A temperature sensor 51 is fitted in an outlet line 50 extending from the coil 16 to the junction 24 with the liquid line 19. The sensor 51 is connected by a tube or lead 52 to a remote sensing proportioning controller 53 which positions a modulating actuator 55 through a lead 54. The sensor 51 is located on a side of the supply coil 16 which serves as an outlet side when the system operates in the heat pump mode, and this sensor generates an actual first reference temperature signal 65. The signal 65 reflects actual refrigerant condensing temperature and is fed into the controller 53 which is pre-programmed with specific parameters. The controller, in turn, generates a first control signal 56 which is outputted to the actuator 55. The control signal 56 represents a difference between the actual first reference temperature 65 and a predetermined first reference temperature as pre-programmed in the controller 53 and is important to the invention as will be described. The proportioning controller 53 is an ambient compensated device, such as a Honeywell T A 991A 1061, having a range from 0 to 135 ohms. The modulating actuator 55 has a rotating arm output complete with an adjustable stroke and adjustable dead band response. The actuator 55 is directly mechanically coupled to face dampers 62 to control flow of air therethrough. The face dampers 62 are coupled through a telescopic coupling means 57 to bypass dampers 58 which are normally gravity-biased to a closed position and are an important feature of the invention. Both dampers 58 and 62 are parallel blade type and serve as variable air flow control means and determine flow through the supply coil as will be described. The actuator 55 can be a Honeywell M 984 which is electrically coupled by a lead 59 to a temperature control system 60 which is described with reference to Figure 12. The system 60 is connected by a lead 63 to a motor 64 controlling the modulating suction cut-off valve 37 as will be described. The face and bypass dampers 62 and 58 and associated couplings will be described in more detail with reference to Figures 2, 3 and 7-11. As previously stated, the direction of refrigerant flows as described above relate to operation of the circuit in the heating mode. When the system is reversed to function as an air-conditioner or refrigeration mode, the reversing valve 34 is rotated through 90° so as to couple the line 36 to the line 38, and the line 21 to the line 40. Clearly, refrigerant vapour passes through the compressor 12, and then passes through the line 21 into the exhaust coil 15, now acting as a condenser, where it leaves as liquid to pass into the junction 24 with the lines 19 and 27. From here liquid returns down the line 27 to the valve 28 as before. Liquid also flows from the junction 24 to the expansion valve 46 which meters liquid refrigerant into the coil 16 where it evaporates and gathers heat. Vapour passes along the line 38 from the supply coil 16 acting as an evaporator, where it returns through suction lines 36 and 42 to the compressor 12. As flow of the refrigerant through the circuit in the air-conditioning mode is identical to the prior art, further description is deemed unnecessary. Figures 2 and 3 As previously stated. Figures 2 and 3 show air flow through the system and damper positions for the heating mode only. Figure 2 showing one hundred percent fresh air heating, and Figure 3 showing optional one hundred percent re-circulation heating. In both fresh air and re-circulation modes, air is moved through a supply chamber 67 of the apparatus 10 under the influence of a supply air fan 68, and through an exhaust chamber 69 by an exhaust air fan 70. The fan 68 is located closely adjacent the supply air coil 16, and draws air therethrough, and the fan 70 is located closely adjacent the exhaust air coil 15 and blows air therethrough. In Figure 2, in the fresh air heating mode, fresh air is drawn from the outside through a pre-filter 71 mounted upstream of the face dampers 62. A horizontal partition 73 divides an inlet air flow path through the chamber 67 into initially separate main and bypass ducts 76 and 77 respectively. The main duct includes the face damper 62, and an inlet space 75 disposed between the face dampers and the supply air coil 16. The bypass duct 77 extends from the pre-filter, through a bypass restrictor screen 79 and then through the bypass dampers 58. The screen 79 is selected to provide a resistance to air flow approximately equal to the resistance to air flow of the coil 16. Thus, when the dampers 58 and 62 are both half open, volume flow through the ducts 76 and 77 is approximately equal. It can be seen that the bypass dampers 58 serve as bypass flow control means cooperating with the bypass duct 77 to control flow of the bypass stream of air therethrough. Similarly, it can be seen that the face dampers 62 serve as main flow control means cooperating with the main duct 76 to control flow of air therethrough.The fan 68 is located so that an intake thereof receives air from both the coil 16 and the bypass dampers 58, if open. .An auxiliary electrical heater 82 receives air from the fan 68 and can provide additional heat to the space to be heated with supply air, as needed. Auxiliary heat is needed when there is insufficient capacity in the heat pump alone to meet the fresh air heating demand as is well known in other prior art heat pump systems. First and second re-circulating dampers 85 and 86 are disposed between the inlet air flow path 74 and a return air flow path 88. The dampers 85 and 86 are also parallel blade type and can be positioned either open or closed only, i.e. no intermediate positions, and are controlled concurrently through a two-position damper operating motor 90. In Figure 2, in the fresh air mode, the motor 9O has been actuated so that the re-circulation dampers 85 and 86 are closed. Barometric dampers 92 are fitted upstream of the fan 70 so as to control air flow into the fan 70, and thus flow through the coil 15 and out to ambient air as exhaust. The recirculation dampers 85 and 86 serve as recirculation flow control means disposed between the inlet chamber and the space being heated. The recirculation flow control means can be set at either open or closed positions as determined by the mode of operation of the equipment. The barometric dampers 92 are also conventional parallel blade type dampers which are gravity biased closed and which automatically open to pass air outwardly from the building when the pressure difference across the damper is greatest on the side of the dampers exposed to room air. In Figure 3, in the re-circulating mode, the motor 90 has been actuated to open the first and second re-circulating dampers 85 and 86 respectively, and the actuator 55 and coupling means 57 have been actuated to close the face and bypass dampers 62 and 58 respectively. The exhaust chamber 69 is now exposed to suction from the supply chamber 67 and thus, because the pressure differences across the dampers 92 is essentially zero, the dampers 92 will automatically close. In the re-circulating mode, return air from the heated space is now drawn through the second re-circulating dampers 86 into the inlet space 75, and back through the coil 16. The fan 68 draws the re-circulated air through the coil 16 to be heated and is then fed back into the space to be heated. The closed dampers 92 prevent mixing of the two streams of air in the inlet and return flow paths 74 and 88 respectively. On the opposite side of the dampers 62 and 92, fresh air is drawn through the re-circulating dampers 85, through the fan 70 and is then forced through the coil 15 back to atmosphere simultaneously being cooled by the coil 15 as the refrigerant boils and extracts heats from outside air as in a conventional heat pump. As previously stated, the telescopic coupling 57 couples the bypass dampers 58 to the face dampers 62 and ultimately to the actuator 55. As will be described with reference to Figures 7-11, the telescopic coupling 57 serves as a means to engage and disengage the coupling between the face dampers 62 and the bypass dampers 58. When the telescopic coupling is engaged, the actuator 55 is directly mechanically coupled to both the face and bypass dampers 62 and 58 so that the dampers operate in equal and opposite directions to each other. That is, when the face damper 62 is fully open, the bypass damper 58 is fully closed and vice versa. This is attained by suitable adjustment of linkage lengths and lever angles as will be described with reference to Figures 9-11. Thus, it can be seen that there is an inverse proportionally between the bypass and face dampers. Referring to Figure 2, the face and bypass dampers are shown to be half open, so that flow is generally equally restricted to both the main duct and the bypass duct. In this mode, the coupling means 57 is engaged and so the dampers 58 and 62 are inversely proportionally coupled. Thus, as the face dampers 62 approach the closed position, the bypass dampers 58 open and this decreases flow through the coil 16, and increases flow to bypass the coil. This action occurs in a situation where fresh air temperature decreases thus requiring a smaller volume flow of colder supply air through the condenser 16. Clearly, the colder the outside air, the less volume is required to maintain optimum balance between the condenser and evaporator. The refrigeration system can now be balanced with precision through a range of temperature far wider than in prior art constant volume fresh air systems. On the other hand, in Figure 3, in the re-circulating mode, the telescopic coupling means 57 is shown disconnected, permitting both the bypass and the face dampers to be closed simultaneously ensuring that return air from the room to be heated passes back into the room after passing through the coil 16. In the reverse mode, i.e. in a one hundred percent fresh air cooling mode, not shown, the telescopic coupling 57 is engaged and the face dampers 62 are fully open, while the bypass dampers 58 are fully closed. In this mode, all the incoming fresh air flows through the evaporator coil to prevent freeze-up. The re-circulating dampers 85 and 86 are closed, and the barometric dampers 92 are fully open. Clearly, the auxiliary heater 82 is disabled, and the modulating suction cut-off valve 37 is fully open. In the fresh air cooling mode the apparatus 10 functions as a conventional refrigeration/air- conditioning system with the exception that the exhaust air coil receives air from the space being cooled, which is usually at a temperature cooler than ambient. Consequently, temperature difference for the coil 15 is greater than normal, and thus heat transfer is improved and "low temperature" that would normally be lost is partially reclaimed. In conventional air conditioning, an auxiliary fan usually drives outside air through the condenser coil, but by using the exhaust fan 70 to drive cooled room air through the condenser coil 15, the auxiliary fan which would otherwise normally be required, is eliminated. This is an extra advantage in the present invention. In summary, less energy is required for the invention as the coil 15 is operating in a lower temperature, which increases heat transfer and refrigerant sub-cooling, and the auxiliary fan is not needed. In the one hundred percent re-circulating cooling mode, not shown, the apparatus 10 functions as a conventional closed loop air-conditioner, with the re-circulating dampers 85 and 86 both opened similar to that as shown in Figure 3. The bypass and face dampers 58 and 62 are also closed as shown in Figure 3 so that return air from the room being cooled passes directly through the re-circulating dampers 86 and the supply coil 16 and back into the room to be cooled. Similarly, fresh air is drawn through the re-circulating dampers 85 into the fan 70 and out through the exhaust air coil 15. Further description of the cooling modes of operation is deemed unnecessary due to the similarity with conventional air-conditioning systems. Figures 4 through 6 The apparatus 10 can be incorporated into a single housing 100 so that the supply and exhaust portions are coupled, thus simplifying some construction. Alternatively, the supply and exhaust portions can be separated and fitted into separate housings (not shown) which are located in different portions of the building. The description following relates to the coupled system which has certain advantages as will be apparent. As seen best in Figure 4, the housing 100 is a generally rectangular structure having a supporting framework, side, top and bottom panels and access doors, not shown, where required to facilitate servicing. The housing 100 is divided internally by a main vertical dividing wall 102 which separates supply and exhaust portions 104 and 105 respectively. As also seen in Figures 5 and 6, the supply and exhaust fans 68 and 70 are mounted in lower and upper portions of the housing 100, so that the housing is also divided horizontally for purposes as will be described. Thus, the exhaust air coil 15 is located in a plane above the supply air coil 16 and displaced laterally therefrom, and the bypass damper 58 is located above and displaced laterally from the face dampers 62. The horizontal partition 73 prevents mixing of the main flow in the inlet chamber 75 and the bypass flow. Thus, the supply portion 104 has the supply chamber 67 having main and bypass ducts for receiving separate streams of air, and the horizontal partition 73 maintains separate main and bypass streams of air. The fan 68 serves as propulsion means for generating air flow through at least one of the ducts. Fresh air enters the housing through optional intake louvres 109 which can be installed to minimize entry of rain or snow into the system. As previously described with reference to Figure 3, and as seen in Figure 5, the inlet space 75 is defined in part by the face damper 62, the supply air coil 16 and the horizontal partition 73. The space 75 also has a side wall formed of the lower re-circulating dampers 86 which are coupled to upper re-circulating dampers 85. As seen in Figure 4, the dampers 85 and 86 are controlled by the motor 90 and are opened when the apparatus functions in the re-circulating mode as previously described. Thus, the vertical dividing wall 102 has an aperture therein which is closed by the re-circulating dampers 85 and 86 when in the normal fresh air heating mode as shown in Figure 2, or in the fresh air cooling mode, not shown, but is opened by the re-circulating dampers when in the heating re-circulating mode as shown in Figure 3, or in the cooling re-circulating mode, not shown. Thus, it can be seen that the recirculating dampers 85 and 86 serve as recirculation flow control means disposed between the supply chamber and the space being heated. For most installations, the recirculation flow control means are settable at either open or closed positions, but in some installations it might be preferable to provide one or more intermediate settings in which the dampers 85 and 86 are partially open to attain a mixture of flow. The recirculation flow control means in effect form a first openable partition, i.e. the first and second re-circulation dampers 85 and 86, between the supply and exhaust chambers 67 and 69 to permit communication therebetween when the recirculating flow control means are open. This is shown in Figure 3, and to control flow it can be seen that the exhaust chamber 69 has a second openable partition, namely the barometric dampers 92 and the face dampers 62. The dampers 92 divide the exhaust chamber 69 into a return portion and an exhaust portion communicating with the space to be heated and atmosphere respectively. The return portion is bouncf -.d by the closed dampers 92 and 62 and air from the space is re-circulated through the dampers 86 and the coil 16. The exhaust portion is also bounded by the closed dampers 92 and 62, and fresh air is drawn through the open dampers 85 and the coil 15. Thus, when the invention is assembled into a single unit as shown, the various dampers serve as a convenient openable partition means for diverting the air between adjacent chambers, depending on the mode of operation of the system. This simplifies considerably ducting and servicing, as the whole unit can be accessed and serviced in a compact space. While the advantages above are related to a coupled system housed in a single housing 100, many of the advantages can also be attained if the supply and exhaust systems are separated. In such a separated arrangement, additional ducting would be required but the two aspects of the invention would function identically. Figures 2. 3. 7 and 8 As seen in Figure 2, the face and bypass dampers 62 and 58 are mechanically connected together with the telescopic coupling means 57, and the dampers 62 are directly coupled to the modulating actuator 55. When engaged, the telescopic coupling means 57 rigidly interconnects the dampers 58 and 62 to operate in the inverse proportional relationship as previously described. As seen in Figure 3, when the coupling means is disengaged, the bypass dampers 58 are de-coupled from the face dampers 62, and are free to close automatically, while the dampers 62 operate in response to actuation of the modulating actuator 55. The telescopic coupling means 57 thus has a provision to disconnect the inverse proportional coupling between the face and bypass dampers which is for use in the re-circulating modes. Referring to Figures 7 and 8, the coupling means 57 comprises first and second rods 121 and 122 which have outer ends coupled by radius arms, not shown, to the main dampers 62 and bypass dampers 58 respectively, not shown in Figures 7 and 8. The rods 121 and 122 have inner ends fitted with a magnet 124 and an armature 125 respectively. The magnet and armature are discs of generally similar size, which are fitted for sliding movement within a tubular body 127. The body has first and second end caps 129 and 130 fitted with aligned bores to receive the first and second rods 121 and 122 respectively and to enclose ends of the body 127. Preferably the magnet 124 is a ceramic permanent ring magnet, and the armature 125 is a steel washer which is attracted to the magnet when in close proximity thereto. Spacing 132 between opposed faces of the magnet and armature equals full stroke of the actuator 55 which is necessary to move the dampers 58 between fully closed and fully opened positions. The body 127 has an axially aligned, elongated clearance slot 133 adjacent the end cap 130 remote from the magnet, the slot extending inwardly from the end cap 130 to an inner end 135 of the slot. A bracket 137 is located adjacent the inner end 135 and journals a latching arm 139, which can slide into the slot as seen in Figures 7 and 8. A transverse pin 138 extends laterally through the arm and interferes with the body 127 to prevent excessive movement of the arm 139 into the body. The arm 139 has an outer end 141 which has mutually perpendicular longitudinal and radial faces 143 and 144, the terms longitudinal and radial referring to the body 127. In Figure 7, the coupling 59 is shown in an extended or disengaged position in which the damper 58 is de-coupled from the actuator 55, as found in the re-circulating mode. In this position, the magnet 124 is adjacent the end cap 129, and the armature 125 is retained adjacent the end cap 130 by the latching arm 139 which has passed into the clearance slot 133 so that the radial face 144 is against the armature 125. A de-coupling stop 150 is located on an adjacent surface 151 to prevent further movement of the means 57 in the direction of the arrow 148. The stop includes a bracket 153 with an adjusting screw 152 which can be adjusted so that the screw 152 contacts the end plate 129 so as to hold the bypass dampers closed when the coupling means 57 disengages the actuator 55 (Figure 3) . Figures 7, and 9 through 11 Figure 9 shows the coupling means 57 in an engaged configuration in which the armature 125 is retained against the magnet 124 so that the rods 121 and 122 are effective axially joined together to function as a single rod. Single blades of the dampers 58 and 62 are shown, the remaining blades being coupled in accordance with common practice. The blade of the bypass dampers 58 has a bypass damper spindle 154 and a bypass damper radius arm or lever 155 extending therefrom. The dampers are positioned so that the spindle is generally horizontal, and blades of the dampers are offset so as to be unbalanced so as to close under gravity, that is, the dampers 58 are gravity biased to automatically close. The blade of the face dampers 62 has a face damper spindle 156 and a face damper radius arm or lever 157 extending from the spindle as shown. The spindle 156 is similarly journalled horizontally, and the face dampers are coupled to an output shaft of the actuator 55 not shown. The first and second rods 121 and 122 are connected to the levers 157 and 155 respectively so that rotation of the spindle 156 by the actuator 55 results in a corresponding rotation of the spindle 154. Thus, when the coupling means 57 is engaged as shown in Figure 9, rotation of the lever 157 in an anti-clockwise direction per an arrow 158 rotates the lever 155 in a clockwise direction per an arrow 159. Referring to Figure 10, for the apparatus to shift from a fresh air configuration to a re-circulating configuration the dampers 58 and 62 must de-couple as follows. The face dampers 62 are rotated anti-clockwise towards a fully closed position as shown by the arrow 158, which in turn draws the rod 121 in direction of the arrow 148 which rotates the bypass dampers 58 anti-clockwise to the fully open condition. The stroke and length of the rods are adjusted so that bypass dampers reach the fully open position before the face dampers are completely closed. Further travel of the rods 121 and 122 to fully close the face dampers exerts an excessive force between the magnet 124 and the armature 125, which force overcomes the magnetic attraction so that the rods eventually separate at a threshold position as seen approximately in Figure 10. Referring to Figure 11, the spindle 156 of the face dampers 62 continues rotation in direction of the arrow 158 until the dampers 62 attain the fully closed position. When the armature 125 is released from the magnet, force from the gravity biased bypass dampers 58 rotates the spindle 154 in the direction of the arrow 158 and draws the rod 122 outwardly through the body 127 in direction of the arrow 145 until the latching arm 139 and the armature 125 assume the position as shown in Figure 7. In this position the bypass dampers are locked closed by both the latching arm 139 and the decoupler stop 150. The coupling means 57 permits either simultaneous inversely proportional movement of the dampers, or closing of the bypass dampers with independent opening and closing of the face dampers. Other means can be substituted to attain the same result, e.g. use of a separate, independently controlled actuator to control the bypass dampers. Thus, the coupling means 57 further includes de-coupling means cooperating with the coupling means to de-couple the main flow control means from the bypass means as required, so as to cancel the inverse proportional relationship between the flow control means. When coupled, the first and second rods are operatively connected together to provide the inverse relationship, and can be separated as required. The coupling means 57 can be retracted by reversing the above procedure, which occurs when the coupling 71 is required to re-couple the face and bypass dampers. As seen in Figure 7, the rod 121 is moved axially inwardly into the body 127 per arrow 145, so that as the magnet 124 approaches the armature 125, the magnet pushes against the latching arm 139, and across the longitudinal face 143 swinging the arm outwardly through the slot in direction of an arrow 146. When the magnet 124 contacts the armature 125, the latching arm 139 is maintained disengaged from the armature 125. Thus, as the rod 121 moves outwardly in direction of the arrow 148, the armature and rod 122 can then move concurrently with the rod and magnet 124, thus moving the entire coupling means to reopen the bypass dampers 58 as described. Thus, in summary, when the armature 125 and the magnet 124 are in contact, the coupling means 57 is coupled or in retracted condition, as shown in Figure 9, and the bypass dampers move in inverse proportion to the face dampers. However, when the coupling means 57 is de-coupled by passing the threshold position as shown in Figure 10, the face dampers become independent of the bypass dampers which remain closed until the coupling means 57 is re-coupled. It can be seen that the body 127 serves as a guide means for guiding extension and retraction movement between the first and second rods 121 and 122. Also, it can be seen that the magnet and armature 124 and 125 respectively serve as magnetic means for coupling together inner ends of the rods. The latching arm 139 serves as a retaining means for retaining one rod in a position while magnetic force coupling the rods together is overcome, permitting the rods to separate and de-couple the flow control means. Figure 12, with some references to Figure 1 The term "signal" herein refers to generation of an electrical characteristic, or fluid pressure or equivalent which is used to actuate or enable a component. The signal can change in magnitude and other characteristics as it passes through several components, but even after such changing it is still referred to as the original signal. For example, an electrical signal when initially generated can be too small, and requires amplification before usage. The numerical designation of such a signal does not change as it passes through the components which can change its characteristic. A portion of a control system 160 is shown, in which the components are shown actuated for the heating mode. A room thermostat 161 is in the air space to be heated and initiates a demand for heat when the actual temperature drops below a pre-set desired temperature commonly termed the thermostat set-point, but defined herein as a predetermined second reference temperature as will be described. The demand for heat initiates a thermostat out- ut signal which has three functions which are considered as three separate thermostat signals from the thermostat 161 as follows. A first thermostat signal 162 is an enabling signal which is fed along the lead 59 (Figure 1) to the actuator 55 to enable the actuator. In response to the actual temperature signal 65 from the sensor 51, the proportioning controller 53, in turn generates the first control signal 56 to the modulating actuator 55 to control the dampers as previously described. It is added that the enabling signal 162 from the thermostat 161 to the actuator 55 and the controller 53 is necessary to prevent operation of the actuator 55 and dampers 58 and 62 during the cooling mode. The room thermostat 161 also generates a second thermostat signal 164 which is fed to a DC amplifier 163 which amplifies the signal and feeds it through a multi-contact relay 165 and to the motor 64 which controls the modulating suction cut-off valve 37. With a full heating demand, the valve 37 is maintained in a fully opened position to permit full flow of refrigerant to the condenser or supply coil 16 (Figure 1) to provide maximum heat output from the apparatus to satisfy the thermostat demand. The second thermostat signal 164 represents difference between an actual second reference temperature (i.e. the actual air temperature) and the desired air temperature or thermostat set point (previously defined as the predetermined second reference temperature) . The difference between the actual second reference temperature and the predetermined second reference temperature is important to the invention and after processing is defined as the second control signal 164. Third and fourth thermostat signals 166 and 168 are generated by the thermostat and fed to an air switch 167 and interlocked thereby. The third signal 166 continues through another air switch 169 and is further interlocked thereby and proceeds to a compressor protection circuit 171. It is further interlocked in the circuit 171 and eventually starts the motor compressor 12 if all interlocks are satisfied. The fourth signal 168 continues from the air switch 167 to energize the auxiliary heater 82 whenever the fourth signal is present and the interlock is satisfied. Thus, the actuator 55 and the motor compressor 12 are effectively configured for simultaneous operation whenever the system is in the heating mode. The diagram shows that the reversing valve 34 does not receive a signal when the thermostat 161 is calling for heat as this is a biased valve which, when non-energized, is biased to the position shown in Figure 1, which as previously stated, discloses the heating mode. The circuit 160 also has a control system 173 for providing demand defrost for the exhaust coil. The system 173 can be a known unit and has a group of components which sense when the exhaust coil 15 starts to accumulate frost, that is accumulate ice on the outside, which restricts flow therethrough. If ice accumulates, the demand defrost initiates a defrost cycle, and it also senses when the defrost cycle should terminate. Defrost logic is passive until required, and is only normally required in the 100% recirculating heating mode. Because the exhaust coil 15, which is acting as an evaporator in the heating mode, will always be operating above 0 β C (32°F) the exhaust air should never be less than 20°C (70°F) in the 100% fresh air heating mode. But, should an exhaust fan air switch fail, the defrost logic will prevent a "snowball" effect resulting in blocking the coil 15 with ice. For changing to the cooling mode, a signal from the air switch 167 is fed to the reversing valve 34 to rotate the valve to the opposite position, not shown in Figure l. This reverses flow in some conduits as described, and the system then operates as a normal air conditioner as will be clear to a person skilled in the art. Clearly, in the cooling mode the modulating suction cut-off valve 37 is fully opened due to a signal from the relays 165, and the auxiliary heater 82 is maintained de-activated by the thermostat 161 as the fourth signal 168 is not present in the cooling mode. OPERATION Referring to Figure 2, in the fresh air heating mode, the fan 68 generates main and bypass streams of air which flow along the main duct 76 and bypass duct 77 respectively, the two streams of air being maintained separate from each other by the partition 73. The main stream of air is directed through the coil 16 which acts as a condenser of vapour compression heat pump and heats the main stream of air. The controller 53 is pre-programmed by the installer to establish the pre-determined base or first reference temperature which is the pre-determined or "ideal" refrigerant condensing temperature. This "ideal" temperature is a compromise between efficiency and comfort and is dependent upon the pressure of the system, type of refrigerant used and other parameters. The sensor 51 is located on the outlet of the condenser coil 16 and thus detects actual refrigerant condensing temperature. The first actual reference temperature 65 outputted by the sensor 51 is fed into the controller 53, compared with the pre-determined first temperature and the difference is outputted from the controller 53 as the first control signal 56, as previously described with reference to Figure 1. This first control signal is fed along the lead 54 to the actuator 55 which, as previously described, controls the dampers. It is noted that the first thermostat signal 162 cannot be generated until the thermostat generates the third signal 166 because power to the actuator 55 is not available until the compressor 12 is energized for use in the heating mode. In response to the first control signal from the controller 53, the actuator 55 positions the telescopic coupling means 57 to control opening of the dampers 58 to control flow of the bypass air in response to the first control signal. Thus, if the first control signal 56 reflects an actual first reference temperature 65 below the predetermined first reference temperature, the bypass flow is increased. Conversely, if the first signal reflects an actual first reference temperature above the predetermined first reference temperature, the bypass flow is decreased. The final step of the heating process is mixing the main stream of air after passing through the condenser 16 with the bypass stream of air, if any air were passed through the bypass dampers 58, prior to delivering the mixture of air streams to the space to be heated through the fan 68. As described previously with reference to Figures 9 through 11, the inverse proportionality between the bypass dampers 58 and face dampers 62 permits controlling the main stream of air also in response to the first signal in an inverse flow ratio to the bypass stream of air. In this way, as flow through the bypass duct increases, flow through the main duct decreases and vice versa while maintaining constant volume flow of air into the space. The second thermostat signal 164 from the thermostat 16, after processing by the amplifier 163, is outputted to the motor 64 to control flow in the suction line 36 by modulating the valve 37. For a normal full demand for heat, the valve 38 is maintained open. As the actual second reference temperature (i.e. actual room temperature) approaches the predetermined second reference temperature, (i.e. the thermostat set point or desired air temperature) the difference between these temperatures diminishes. This difference is the second control signal 164, and the condenser heat rejection is reduced in proportion to this diminishing control signal by inhibiting the ability of the evaporator to gather heat from air leaving the building in the exhaust stream. Thus, the evaporator operates at an artificially high pressure while the compressor (which is exposed to suction of the modulating cut-off valve 37) operates at a lower than normal pressure, that is the compressor is starved of refrigerant. Volumetric efficiency of the compressor is diminished in inverse proportion to the compression ratio which further reduces system capacity. The de-superheating expansion valve 28 prevents compressor overheating while operating under these conditions for sustained periods by admitting a small amount of refrigerant into the compressor as required. The valve 37 is automatically positioned in response to the second control signal 164 from the thermostat 161 reflecting heating demand, and as this demand decreases, the valve is gradually shifted proportionally towards the minimum flow position. Thus, the valve is modulated so as to balance the condenser heat rejection to the fresh-air heating load, thus maintaining a constant supply air temperature throughout throttling range of the valve 37. As previously stated, the higher internal evaporator pressure caused by the throttling action of the valve 37 results in a higher boiling temperature of the refrigerant passing through the evaporator. This reduces the net temperature difference between the evaporator air stream which decreases the heat gathered by the evaporator which in turn decreases the condenser demand or heat rejection. As the total heat rejection at the condenser is equal to the heat gathered in the evaporator plus the heat of compression, it follows that heat rejected at the condenser, termed condenser demand, can be controlled by controlling the amount of heat gathered in the evaporator. The said modulation of the valve 37, which gradually reduces refrigerant flow, prevents cold ambient air from entering the building during continuous fresh-air supply. Use of the valve reduces considerably the off-cycle periods of the refrigeration plant when the refrigeration plant output would normally exceed the fresh-air heating demand. It can be seen that the valve 37 effectively provides capacity control of a heat pump without a multi-stage of multi-speed compressor, and thus provides a considerable simplification over control systems of prior art heat pumps. Thus, when the air in the room actually attains the thermostat set-point, that is the second control temperature 164 signal is 0, capacity of the heat pump has been reduced considerably by the modulating valve 37, and thus temperature of air passing into the room essentially equals temperature of air already in the room. Thus, the heat pump maintains temperature in the room between "on and off" cycles as controlled by the thermostat without surges of cold air which would otherwise be generated when the refrigeration plant output is greater than the building heating demand. Usually, a normal heating demand signal from the thermostat 161 positions the valve 37 to a fully open condition so that heating output is at a maximum. When the compressor initially starts to pressurize refrigerant, in response to the initial full demand of the second control signal from the thermostat, temperature at the sensing bulb 44 is initially increased. This will cause the de-superheating thermostatic expansion valve 28 to admit liquid refrigerant from the liquid line 27 into the suction line 42, via the pressure vessel 30. This lowers the compressor discharge temperature, and high pressure heated vapour passes through the reversing valve 34 into the vapour line 38 towards the supply coil 16. In the coil 16, the refrigerant condenses, passing heat into the supply air, and liquified refrigerant bypasses the expansion valve 46 through check valve 48 to arrive at the junction 24. Clearly, the temperature sensor 51 detects the temperature of this liquid refrigerant, which in conjunction with the proportioning controller 53, positions the dampers as required. It is added that if the condensing temperature at the sensor 51 becomes too low, the proportioning controller 53 re-positions the face dampers to reduce flow through the coil 16, and, through the ϊrse proportional coupling 57, opens the bypass dampers 58 to a wider position to maintain constant air volume to the supply fan 68. In effect, the air quantity through the supply coil 16 is modulated by the face dampers to attain optimum refrigerant condensing temperature. It is added that the thermostatic expansion valve 23 on the exhaust coil 15 functions in a normal manner, in response to outlet temperature from the exhaust coil as detected by temperature sensor 26. When the liquid refrigerant passes through the expansion valve 23, the pressure in the exhaust coil 15 drops so that the refrigerant boils and the resulting vapour gains heat from air drawn through the exhaust coil by the fan 70. The resulting vapour is slightly super-heated and is drawn towards the compressor through the reversing valve 34 and then to the modulating suction cut-off valve 37. As stated previously, the cut-off valve 37 controls the compressor compression ratio and thus the heat gathered and rejected, as well power consumed by the compressor. Gas drawn through the valve 37 passes into the pressure vessel 30 which traps liquid refrigerant to avoid being drawn into the compressor. The compressor is suction cooled, but if the compressor begins to overheat, the de-superheating expansion valve 28, controlled by the sensor 44, will admit sufficient liquid refrigerant into the suction line to lower the gas superheat to cool the compressor. Usually, the only time this would occur would be during prolonged periods of compressor operation at minimum load. When a signal for cooling is initiated from the thermostat 161, the reversing valve 34 is energized, and the relay 165 is activated to shift the modulating cut-off valve 37 to a fully open position. The relay 165 also overrides the controller 53 and causes the actuator 55 to move the bypass dampers 58 to a fully closed position, which simultaneously results in the face dampers 62 being moved to a fully open position. The compressor 12 also responds to the thermostat demand for cooling and must satisfy the same interlocks as is required in the heating mode compressor operation. The remaining portion of the refrigeration and cooling cycle is similar to the prior art, and consequently, further discussion is deemed unnecessary. ALTERNATIVES The apparatus as disclosed receives two main control signals which initiate and control the heating sequence and operation of components. The first control signal 56 is generated by the controller 53 (in response to the signal 65 from the sensor 51) and reflects difference between actual and pre-determined refrigerant temperatures. This first control signal, via the modulating actuator 55, controls the face and bypass dampers to control temperature, and thus pressure of the refrigerant in the supply coil. The sensor 51 would normally be positioned in a location to detect actual refrigerant condensing temperature and this location follows normal practice and is tne preferred location of the sensor 51, but temperature at other locations in the supply coil could also be used. As air leaving the supply coil reflects, to some extent, refrigerant temperature in the supply coil, it could be possible, for some applications, to locate the sensor 51 directly in the air stream from the supply coil. Clearly, in this location, the thermostat should be relocated elsewhere. In some applications, it might be possible to eliminate one temperature sensor, and to provide a combined signal which could be suitably adjusted to actuate both the actuator 55 and the modulating valve 37. The invention discloses variable face dampers 62 which control the main flow of air through the supply coil 16 in inverse proportion to flow through the bypass dampers, thus maintaining essentially constant total air flow into the air space. In some applications, the variable face dampers could be eliminated, and flow through the supply coil could be effectively controlled remotely by the bypass dampers 58, in a ratio dependent on the relative resistance between flow through the bypass dampers, and flow through the supply coil. It is considered that this alternative would be considerably less sensitive and more difficult to control than the preferred embodiment as disclosed as the total air flow rates would vary depending on the position of the bypass dampers. Also, the modulating valve 37 is provided to permit greater control of operation of the compressor, and thus greater control of heat recovered and heat rejected. In some applications in which the heating demand is greater than the heating capacity, the valve 37 could be eliminated, permitting the compressor to function without this additional control. However, for normal temperate climates where occasionally only a small function of system heating capacity is required, use of the modulating suction cut-off valve 37 is preferred as there are less temperature fluctuations in the discharge air as the thermostat set-point is gradually attained. The invention is disclosed showing an essentially constant volume flow through the supply chamber. This constant flow is attained by proportioning volume flow between the main duct and the bypass duct by controlling the respective main dampers and bypass dampers in an inversely proportional relationship. This is the preferred combination for most installations, and energy efficiency thereof is further improved by providing the modulating suction cut-off valve 37 which reduces heat rejection at the condenser. For some applications, where multiple or unloading compressors are not used, use of the modulating suction cut-off valve 37 as a means of capacity control for a heat pump can be employed with a considerable saving in cost, and in some instances, improvements in efficiency. In some alternate installations, where constant air volume flow is not of importance, the bypass flow duct and associated modulating bypass dampers 58 can be eliminated, and all air flow then passes through a single duct. In this alternative, capacity control using the cut-off valve 37 is sufficient to provide an improvement over prior art methods. In other installations, the telescopic coupling means 57 coupling the dampers 58 to the dampers 62 can be eliminated where one hundred percent recirculation and fresh air options are required. In this instance, the actuator 55 could be directly coupled to the bypass dampers 58, and an alternative means to permit simultaneous closing of both the face and bypass dampers would be required. Also, in some installations where one hundred percent recirculation is desired, the coupling 57 could be eliminated, and the recirculation dampers 85 and 86 could also be eliminated. When one hundred percent recirculation is not required, the actuator 55 could be directly coupled to the bypass dampers 58 by way of a solid rod, thus eliminating the option for uncoupling the bypass dampers. The temperature sensors cooperating with conduits carrying refrigerant are expanding bulb types connected with capillary tubes to valves or other control means. While this is the preferred means of detecting temperature of the refrigerant, clearly equivalent temperature sensors could be substituted, for example sensors which directly generate electrical signals which are then used to actuate valves or related controls. Also, the thermostat of the control system as described is a preferred currently available unit which generates a plurality of signals for actuating or enabling other components of the system. Clearly, alternative control systems with similar functions can be substituted. The exhaust air coil 15 is disclosed for location in the exhaust chamber 69, and thus receives a tempered stream of inside air exhausted from the room. While this location of the exhaust coil is preferred because some room heat can be recovered from the exhaust air in the heating mode, and improved cooling generally results in the cooling mode, alternative locations of the exhaust coil are envisaged. Clearly, the exhaust coil can be located to be exposed to ambient or outside air as in a conventional air conditioning system, and this might be appropriate in certain installations. Also, heat could be gathered by the exhaust coil from waste water or a ground heating source. Also, as stated previously, a mixture of recirculated inside air and fresh air can be attained by stopping the motor 90 in an intermediate position, so that the recirculation dampers 85 and 86 are partially open, permitting some mixing of fresh air with the inside room air.

Claims: CLAIMS : 1. A method of heating air comprising the steps of: (a) generating main and bypass streams of air, and initially maintaining the two streams of air separate from each other, (b) directing the main stream of air through a condenser of a vapour compression heat pump to heat the main stream of air, (c) generating a first control signal representing a difference between an actual first reference temperature and a pre-determined first reference temperature, (d) controlling flow of the bypass stream of air in response to the first control signal so that if the first control signal reflects an actual first reference temperature below the pre-determined first reference temperature, the bypass flow is increased, and if the first signal reflects an actual first reference temperature above the pre-determined first reference temperature, the bypass flow is decreased, (e) mixing the main stream of air after passing through the condenser with the bypass stream of air, if any air were passed through the bypass flow control means, prior to delivering the mixture of air streams as required to a space. 2. A method as claimed in Claim 1, further comprising: (a) controlling the main stream of air also in response to the *:irst control signal in an inverse flow ratio to the bypass stream of air, so that as flow through the bypass duct increases, flow through the main duct decreases and vice versa. 3. A method as claimed in Claim 1, further comprising: (a) generating a second control signal reflecting difference between an actual second reference temperature and a pre-determined second reference temperature, (b) gradually reducing heat rejection at the condenser as the actual second reference temperature approaches the pre-determined second reference temperature. 4. A method as claimed in Claim 3, in which: (a) reducing the heat rejection at the condenser is attained by reducing heat gathered in an evaporator of the heat pump from air passing through the evaporator. A method as claimed in Claim 1, further comprising: (a) generating a second control signal representing a difference between an actual second reference temperature and a pre-determined second reference temperature, (b) gradually reducing flow of refrigerant to a compressor of the heat pump as the actual second reference temperature approaches the pre-determined second reference temperature so that volumetric efficiency of the compressor is reduced and capacity of an evaporator of the heat pump is reduced. 6. A method as claimed in Claim 1, wherein the first reference temperature reflects temperature of air delivered and mixed in the space to be heated. 7. A method as claimed in Claim 2, wherein the second reference temperature reflects temperature of refrigerant in the condenser coil. 8. A method as claimed in Claim 1, further characterized by recovering some heat from air being exhausted from the space to be heated. 9. A method as claimed in Claim 9, further characterized by: (a) locating an evaporator of the heat pump in the air flow exhausted from the space. 10. An apparatus for heating air comprising: (a) a supply chamber having main and bypass ducts for receiving and initially maintaining separate main and bypass streams of air, (b) propulsion means for generating air flow through at least one of the ducts, (c) a condenser of a vapour compression refrigeration plant provided in the main duct to heat the main stream of air passing along the main duct, (d) first temperature sensor means for generating a first control signal representing difference between an actual first reference temperature and a pre-determined first reference temperature, (e) bypass flow control means for controlling flow of the bypass stream of air through the bypass duct, the bypass control means cooperating with the bypass duct, (f) actuator means for actuating the bypass flow control means so as to vary an amount of 5 restriction to air flow through the bypass duct, the actuator means being operatively coupled to the bypass flow control means and being responsive to the first control signal, 10 (g) mixing means for mixing outlet flows discharged from the main and bypass ducts prior to discharging the mixed streams of air into the space to be heated. •15 11. An apparatus as claimed in Claim 10, further comprising: (a) main flow control means cooperating with the main duct to control flow therethrough, 20 (b) coupling means coupled to the main flow control means and the bypass flow control means so that the main flow control means is actuated in an inverse relationship to the bypass flow control 25 means. 12. An apparatus as claimed in Claim 10, further comprising: 30 (a) second temperature sensor means for generating a second control signal representing difference between an actual second reference temperature and a pre-determined second reference temperature, 35 (b) heat rejection reducing means for gradually reducing heat rejected at the condenser, the heat rejection reducing means cooperating with the condenser so that as the actual second reference temperature approaches the pre-determined second reference temperature, heat rejected to air passing through the condenser is gradually reduced. 13. An apparatus as claimed in Claim 12, in which: (a) the heat rejection reducing means is a fluid flow modulating valve cooperating with a conduit communicating with an intake of a compressor of the heat pump, the fluid flow modulating valve being responsive to the second temperature sensor so that as the actual second reference temperature approaches the pre-determined second reference temperature, flow of refrigerant through the condenser is reduced so that heat rejected at the condenser is reduced. 14. An apparatus as claimed in Claim 11, further including: (a) the main flow control means being located in the main duct and spaced upstream from the condenser to define an inlet space therebetween with the main duct, (b) re-circulation flow control means disposed between the inlet space and the space being heated, the re-circulation flow control means being settable at or between open or closed positions. 15. A apparatus as claimed in Claim 14, in which: (a) an exhaust chamber communicates with the space being heated to receive air being exhausted therefrom, (b) the recirculating flow control means forms a first openable partition between the supply and exhaust chambers to permit communication therebetween when the recirculating flow control means are open. 16. An apparatus as claimed in Claim 15, in which: (a) the exhaust chamber has a second openable partition to divide the exhaust chamber into a return portion and an exhaust portion communicating with the space to be heated and atmosphere respectively, (b) the main flow control means can be closed together with the second openable partition when both recirculation dampers are closed in the recirculating mode to direct return air from the space back into the space. 17. An apparatus as claimed in Claim 11, further including: (a) de-coupling means cooperating with the coupling means to de-couple the main flow control means from the bypass means as required, so as to cancel the inverse relationship between the flow control means. 18. An apparatus as claimed in Claim 17, in which: (a) the coupling means includes first and second rods operatively connecting together the bypass flow control means and the main flow control means to provide said inverse relationship, (d) the de-coupling means is operative to separate the rods when required. 19. An apparatus as claimed in Claim 18, in which the coupling means comprises: (a) guide means for guiding extension and retraction movement between the first and second rods, (b) magnetic means for coupling together inner ends of the rods, (c) retaining means for retaining one rod in a position while magnetic force coupling the rods together is overcome, permitting the rods to separate and de-couple the flow control means. 20. An apparatus as claimed in Claim 17, in which: (a) the bypass flow control means is releasably coupled to the main flow control means to permit de-coupling therefrom, (b) the bypass flow control means are normally biased to a closed position so that when de-coupled from the main flow control means, the bypass flow control means automatically close. 21. An apparatus as claimed in Claim 10, in which: (a) the heat exchange coil is located in an exhaust portion communicating with the space. 22. A method of heating air comprising the steps of: (a) directing a stream of air through a condenser of a vapour compression heat pump to heat the stream of air, (b) generating a control signal representing difference between an actual reference temperature and a pre-determined reference temperature, (c) gradually reducing heat rejection at the condenser as the actual reference temperature approaches the pre-determined reference temperature. 23. A method as claimed in Claim 22, in which: (a) reducing the heat rejection at the condenser is attained by reducing heat gathered in an evaporator of the heat pump from air passing through the evaporator. 24. A method as claimed in Claim 23, in which: (a) reducing heat gathered in the evaporator is attained by gradually reducing flow of refrigerant to a compressor of the heat pump as the actual reference temperature approaches the pre-determined reference temperature so that volumetric efficiency of the compressor is reduced and capacity of the evaporator of the heat pump is reduced. 25. An apparatus for heating air, comprising: (a) a supply chamber having a duct to receive air to be heated and a propulsion means for generating flow of a stream of air through the duct, (b) a condenser of a vapour compression heat pump provided in the duct to heat the stream of air passing along the duct, (c) a temperature sensor for generating a control signal representing difference between an actual reference temperature and a pre-determined reference temperature, (d) heat rejection reducing means to gradually reduce heat rejected at the condenser, the heat rejection reducing means cooperating with the condenser so that as the actual second reference temperature approaches the pre-determined reference temperature, heat rejected to air passing through the condenser is gradually reduced. 26. An apparatus as claimed in Claim 25, in which: (a) the heat rejection reducing means is a fluid flow modulating valve cooperating with a conduit communicating with an intake of a compressor of the heat pump, the fluid flow modulating valve being responsive to the temperature sensor so that as the actual second reference temperature approaches the pre-determined reference temperature, flow of refrigerant through the condenser is reduced so that heat rejected at the condenser is reduced. 27. A method of heating an air space comprising the steps of: (a) drawing air through a condenser of a heat pump so that condensation of the refrigerant therein heats the air prior to discharging the air into the space, (b) drawing air from the space to be heated through an evaporator of the heat pump to heat the refrigerant, (c) transferring the heated refrigerant thereof from the evaporator to the condenser. 28. A heat pump for heating an air space, comprising: (a) an inlet duct and an outlet duct communicating with the space, (b) means to move the air through the inlet duct into the space and to discharge air from the space through the outlet duct, (c) a refrigerant condenser in the inlet duct and a refrigerant evaporator in the outlet duct, (d) refrigerant conduit means and a compressor communicating the evaporator and the condenser and transmitting refrigerant therebetween to function as a heat pump. 29. A method of cooling an air space comprising the steps of: (a) drawing air through an evaporator of an refrigeration plant, so that evaporation of refrigerant cools the air prior to discharging the air into the space, (b) drawing air from the space to be cooled through a condenser of the refrigeration plant to cool the refrigerant, (c) transferring the cooled refrigerant from the condenser to the evaporator. 30. An air-conditioner for cooling an air space comprising: (a) an inlet duct and an outlet duct communicating with the space, (b) means to move air through the inlet duct into the space, and to discharge air from the space through the outlet duct, (c) a refrigerant evaporator in the inlet duct and a refrigerant condenser in the outlet duct, (d) refrigerant conduit means and a compressor communicating the evaporator and condenser and transmitting refrigerant therebetween to function as an refrigerant plant.

Inventors: DICK BRIAN GEORGE, DICK, BRIAN, GEORGE
Applicants: ALBERNI THERMODYNAMICS LTD, ALBERNI THERMODYNAMICS LTD.
Application year: 1991
Main classification: F24F11
IPC codes: F24F1, F24F11, F25B13