Dynamic outside air management system and method

Abstract

An outside air management system in which the air is cooled by a first refrigeration condenser to a temperature either above or somewhat below the dewpoint of the incoming air so as to remove either none of the moisture to be removed, or a significant portion of it. The air then is sent to a desiccant wheel which removes the remainder of the moisture to be extracted. The temperature of the air leaving the wheel is above the desired delivery temperature. Then a heat exchanger and a second condenser remove sensible heat from the heated air leaving the desiccant wheel to bring the air to the desired temperature. Thus, the temperature and humidity are controlled substantially independently of one another. A heat pump embodiment of the invention allows the reversal of the flow of refrigerant in the system, allowing it to be used to either heat or cool.

Description

Priority for this patent application is claimed from U.S. provisional patent application Ser. No. 61/210,921 filed Mar. 24, 2009. The disclosure and claims of that patent application hereby are incorporated herein by reference.

This invention relates to air conditioning systems and methods. In particular, the invention relates to dynamic outside air management systems and methods, in which the humidity, the temperature and the flow rate of outside air introduced into a building are controlled.

A very common prior art approach to reducing the humidity and/or temperature of mixed outside and return air is one in which refrigeration equipment is used to cool that air to a temperature determined by the apparatus dewpoint which is well below that needed to bring the temperature down to the level desired in the conditioned space. This is done in order to remove enough humidity from the air to meet the humidity requirements for air used in the conditioned space.

Because the temperature of the refrigerated supply air is well below that needed to cool the conditioned space to the desired temperature and humidity, the air may have to be reheated before delivering it to the conditioned space, at part cooling load, e.g.

Such a prior system has at least two major drawbacks; first, that the capital expenditures for the equipment needed for a new HVAC system are relatively high, and, secondly, the energy usage for both a new and a retrofitted existing system is excessive. These disadvantages increase the capital equipment and operating costs for using the system.

In other prior dehumidification systems, desiccant devices are used to assist in dehumidifying outside air. In such systems, it is difficult to control the temperature of the air independently of the desired dewpoint. This often results in the delivery of air which is wrong in one or both of those parameters; that is, it often is either the wrong temperature or wrong dewpoint. Additional equipment suggested to correct the problem is undesirably complex and of limited utility.

Accordingly, it is an object of the invention to provide air conditioning equipment and methods in which the foregoing disadvantages are eliminated or reduced.

It is a further object of the invention to provide a dynamic outside air management system and method in which dehumidification and cooling of outside air can be controlled substantially independently of one another without undue complexity and cost.

In addition, it is an object to provide such a system and method which is capable of delivering outside air in a “space neutral” condition in both heating and cooling operating modes with good efficiency and lower levels of overall HVAC system energy usage.

It is another object to provide a dynamic outside air system and method which can be used for humidification as well as for dehumidification, and for heating as well as cooling outside air and also can perform in a free cooling mode in which either dehumidification or cooling or both can be suspended while providing adequate ventilation for a conditioned space, thereby conserving energy which otherwise might be wasted.

It is another object of the invention to provide automatic reduction of the quantity of outside air when it is not needed for ventilation because of the reduced presence of people in the conditioned space.

In accordance with the present invention, the foregoing objects are met by the provision of an outside air management system and method in which incoming outside air is first cooled to a first temperature and then dehumidified by a desiccant device to a predetermined desired dewpoint, followed by cooling the dehumidified air to a space-neutral temperature, and delivering that air to a utilization location. Thus, the dehumidification is performed by the combination of the first cooling means and the desiccant wheel, and the delivery temperature of the air controlled independently by a separately controllable sensible cooling step.

Preferably, the first temperature is set at a value as high as possible, so as to use as little energy as possible in cooling, while minimizing the energy used in regenerating the dessicant in the desiccant wheel.

Thus, the cost of the equipment is limited, and the temperature of the air leaving the desiccant device can be held at or above the desired delivery air temperature so that further sensible cooling is all that is needed and heating is not required.

Preferably, the system and method operates in three different modes; cooling mode, when the outside air is hot and usually humid; heating mode, when the outdoor air temperature and humidity are relatively low; and “free cooling” mode when one or more of the dehumidification and cooling functions is disabled because little or no heating or cooling is unnecessary.

The operation of the equipment preferably is controlled by a programmable logic controller (PLC) which senses dry-bulb and wet-bulb temperatures of the outside air, flow rate of the outside air, the output of the thermostat in the conditioned space, as well as dew point or wet-bulb/dry-bulb temperatures in the controlled space. The PLC controls the desiccant wheel speed and the variable-output refrigerant compressor to vary the amount of moisture removed from the air, and controls the refrigeration equipment to control the amount of temperature reduction it produces. The PLC controls the fans for outside air and exhaust air during the different modes of operation, and in response to the sensing of CO₂ in the conditioned space to reduce the energy wastage when ventilation can be reduced without harm to the occupants.

The system can be configured as a heat pump system, including reversing valves for converting the system from a cooling mode to a heating mode in which humidification is provided by operation of the desiccant wheel to transfer moisture.

The foregoing and other objects and advantages of the invention will be set forth in or apparent from the following description and drawings.

IN THE DRAWINGS

FIG. 1 is a schematic diagram of a dedicated outside air conditioning system in accordance with the present invention;

FIG. 2 is a schematic diagram of a modified system like that shown in FIG. 1;

FIG. 3 is a schematic diagram of a programmable logic controller used in controlling the systems shown in FIGS. 1 and 2 and elsewhere in this patent application;

FIG. 4 is a schematic diagram of another embodiment of the invention like that of FIG. 1;

FIG. 5 is a psychometric diagram used in explaining the invention;

FIG. 6 is a schematic diagram of a refrigeration system used with the system of FIG. 4;

FIG. 7 is a temperature-entropy plot used in explaining the system of FIG. 4;

FIG. 8 is a schematic diagram of a heat-pump system like that of FIG. 6, in a cooling mode of operation;

FIG. 9 is a schematic diagram of the heat-pump system of FIG. 8 in a heating mode of operation.

FIG. 10 is a schematic diagram of the heat pump system as in FIG. 9, but with mechanical components added as in FIGS. 1 and 2.

FIG. 11 shows an improved version of the FIGS. 9 and 10 system; and

FIG. 12 is a schematic diagram of the FIG. 8 system, with modifications as in FIGS. 10 and 11.

GENERAL DESCRIPTION

FIG. 1 shows a system 20 in which air from outside of a building (“OSA”) is drawn into a building at inlet 22 by a fan 34, processed by the system 20, and delivered at an outlet 36 to a utilization location (“UL”) 24.

The utilization location 24 can be an indoor building space to be conditioned, or the junction with another conditioning system which is used together with the dynamically controlled outside air system 20.

The system 20 preferably has three possible modes of operation, depending upon the climate at the location of the building in which the air is used.

The three modes of operation are:

1. Cooling mode

2. Free cooling mode

3. Heating mode

In some climates, only one of the three modes of operation maybe used. For example, in climates which are constantly hot and humid only the cooling mode might be necessary. In more temperate zones, all three modes may be needed, each at different time of the year.

Certain buildings may have internal heat and humidity loads which render its conditioning needs substantially independent from the external climactic environment. For example, computer data centers which are located around the world develop so much internal heat from the computing equipment and lights, in addition to the heat and humidity generated by people working inside the building, that such buildings may need cooling and dehumidification almost all the time.

Office buildings in which a substantial number of people work and office equipment and lighting are used have similar needs.

Cooling Mode

FIG. 1 shows the system 20 as it is configured for operation in the cooling mode.

The outside air is delivered to a first set of evaporator coils DX2 at 26. The first evaporator DX2 is set to cool the incoming outside air to a temperature which is substantially higher than that used in the ordinary system in which refrigeration is used entirely for dehumidification. The air preferably is cooled only to a temperature sufficiently low to advantageously use it with a desiccant device such as a desiccant wheel, and the desiccant device is used to remove the majority of the moisture to be removed from the outside air. Therefore, the first evaporator DX2 is set to reduce the temperature only as low as needed to avoid or minimize supplemental heating of the desiccant wheel regeneration air. Thus the first temperature will be set at a level which may produce from zero to a significant amount of the moisture to be removed.

The outside air leaves the evaporator 26 and flows to a desiccant wheel 28 which removes the remainder of the moisture to be removed. Thus, the desiccant wheel removes moisture until the air has the desired humidity. Then, the air then flows through a heat exchanger 30 such as a sensible heat wheel, heat pipes, etc. This cools the air leaving the desiccant wheel which has been heated in the process of drying the air.

The somewhat cooled air from the sensible wheel 30 then is delivered to a second set of evaporator coils DX1 at 32 which further sensibly cools the air received from the sensible wheel to a desired temperature for use at the utilization location 24. Because the desiccant wheel is operated so as to deliver air at temperatures at or above the desired temperature, only sensible cooling is needed; heating is unnecessary.

If, for example, the utilization location is a space to be conditioned, the air delivered at the outlet 36 preferably will be at a “space-neutral” temperature and humidity; that is, at the temperature and humidity desired for the air to enter the conditioned space.

Alternatively, at the utilization location 24, the air delivered at 36 can be mixed with return air conditioned by the main refrigeration system, or otherwise, as desired.

It is notable that the air delivered at 36 is delivered at the desired temperature and humidity without requiring reheating, thus saving substantially in equipment and operating costs compared with prior systems using only refrigeration for dehumidification.

Also included in the system 20 is a return duct 38 which delivers return air to one side of the sensible wheel 30 to heat the return air for use in regeneration of the desiccant wheel 28.

Heated air leaves the wheel 30 and moves to a desuperheater 40 which heats the air further.

Next, the air travels to a supplemental heater 42 which can be used if and when necessary to further heat the air delivered to the desiccant wheel 28 to regenerate the desiccant. The air pulled through the return system by the exhaust fan 44 then exits the system at 46.

Thus, heat extracted from the refrigeration system and the desiccant wheel is used to regenerate the desiccant, and make maximum utilization of the energy available in the system.

Optionally, return air can be added to the outside air, as indicated by the dashed line RA, in the amount of 10% to 15%, to somewhat reduce the amount of dehumidification required.

Refrigeration System

The refrigeration subsystem is shown in the lower portion of FIG. 1. It includes a variable output compressor 48, a condenser 50, a subcooler 52 and a single expansion valve 54. In the system shown in FIG. 1, the two evaporators, DX1 and DX2 are connected and parallel to one another and are both controlled by the single expansion valve 54. This is possible where both DX1 and DX2 operate in the same temperature range. The subcooler 52, as it is well known, exchanges heat between the heated and cooled gases and liquids flowing into the subcooler.

The advantageously smaller than conventional condenser uses outside air introduced at 22 for cooling its coils with the aid of a fan 56.

The desuperheater 40, as it is well known, receives hot compressed gas from the compressor 48, uses the heat to heat the return air flowing through it, and sends the gas to the condenser 50.

The system shown in FIG. 1 thus operates in a highly efficient manner whereby the system operates so as to control the dewpoint of the air discharged and its temperature.

FIG. 2 shows a dynamic outside air management system 60 which is the same as the system shown in FIG. 1 except that a modified refrigeration subsystem 23 is provided.

The subsystem 23 is the same as subsystem 21 shown in FIG. 2 except that it has a separate expansion valve 58 for the evaporator DX1, in addition to the expansion valve 54, and the evaporators DX1 and DX2 are not connected to one another in parallel. Thus, the two evaporators can be operated independently of one another, and in different temperature ranges.

Control System and Method

FIG. 3 is a schematic circuit diagram showing a programmable logic controller (“PLC”) 72, together with the sensors from which it receives information, and the operating devices which it controls.

The PLC 72 can be any one of a number of commercially available programmable logic devices which can be programmed, within the skill of the art, to perform the functions to be described below.

The sensors are shown in two separate groups, group 74 and group 76. The sensors in group 74 usually will be located so as to detect the characteristics of the incoming outside air.

The sensors in group 76 normally are positioned in or near the conditioned space or elsewhere where convenient in the building.

The outside air sensors include a flow sensor 78 detecting the flow rate of the incoming outside air, a dry-bulb thermometer 80 and a wet-bulb thermometer 82 to detect the dry-bulb and wet-bulb temperatures of the incoming outside air.

The detectors 76 include a conventional space or return duct mounted thermostat 84, a CO₂ sensor 92, and a dewpoint detector 94. If desired, the dewpoint detector 94 can be replaced by a dry bulb and a wet bulb sensor, such as the thermometers 80 and 82. In the case in which dry bulb and wet bulb temperatures are sensed by separate units and delivered to the PLC, the PLC is programmed to compute the dew point and control additional equipment as described below.

The equipment controlled by the PLC includes the outside air fan 34, the exhaust fan 44, the supplemental heater 42, the speed of the desiccant wheel motor 88 and the speed of motor 86 of the compressor to control its output.

Also, the valves of the refrigeration system, indicated collectively at 90, also are controlled by the PLC.

Further, the PLC controls the main air handling system 99, to control the desired temperature in that space, as it will be described below.

It should be understood that when the system of FIG. 1 or FIG. 2 operate in a heat pump mode, either for cooling or heating, the additional valves needed for such operation are included within the group indicated at 90.

In addition, the sensible heat wheel drive motor 96 can be controlled, if needed or desired, so as to increase or decrease the sensible wheel speed if and when necessary.

An additional device 98 supplies calendar and clock information. Actually, such information normally would be provided internally within the PLC, but it is shown as a separate input, for the sake of explanation.

By using the calendar and clock inputs, the PLC can set the system 20 or 60 for operation in accordance with the time of year and time of day, and whether the building zone occupied or not, thereby conditioning the system for use in various different seasons of the year, and times of day, so as to reduce the unnecessary usage of energy when few or no people are occupying the buildings, and for selecting among the three different operational modes of the system.

Similarly, by use of the CO₂ sensor 92, the ventilation needs of people in the thermostatically controlled zones of the building at a particular time can be determined. This can allow the outside air flow as detected by the flow detector 78 to be reduced significantly by reducing the speeds of fans 36 and 44 because the CO₂ detector detects the reduced CO₂ output by reduced numbers of people. This can reduce the need for ventilating air when demand is very low, thus further saving energy and/or improve operation of separate secondary systems as described below. In addition, energy is saved in the main air conditioning system in an existing building, which uses refrigeration for dehumidification, because the temperature at which the air must be maintained can be raised (e.g., from 55° F. to 57° F.) because there is less outside air entering, and more recirculated air which needs less cooling and dehumidification.

EXAMPLE

Assume a 1400 cfm conventionally designed AHU (air handling unit) zone cooling requirement at a 55° F. supply air temperature and 25% OSA (or 350 cfm) for occupant ventilation needs. Therefore, at zone design conditions, [1400−350] or 1050 cfm of space would by mixed with 350 cfm OSA return air upstream of the AHU cooling coil, discharging supply air at 55° F. Assuming an apparatus dewpoint of 52° F. for a conventional AHU cooling coil with a 55° F. leaving air temperature requiring a constant 25% OSA or 350 OSA to satisfy peak zone occupant demands, the AHU should provide both sensible and latent cooling to maintain a conditioned building space at 75° F. and 50% RH.

Assume the combined cooling load with OSA at peak design cooling load is 58,590 BTU/hr computed as follows: 1400×4.5×9.3 BTU/#BDA (delta h). Building occupants tend to move around during a typical work day, as part of their normal activities, particularly where current “hotelling” company practices allow more of their employees to work at home and come into their offices only when necessary. Therefore, a more realistic OSA CFM rate=0.7×350=245 cfm would seem advisable, particularly when the present dynamically managed pre-conditioner is employed at design cooling day conditions. Accordingly, the corresponding recirculated return air flow rate equals [1400−245] or 1155 cfm.

Assume that either a new or existing conventional AHU is required to provide sensible cooling only. Therefore, its leaving coil temperature can be reset by the programmed PLC and OSA flow rate sensor located in the room return air duct could be automatically reset to 57° F. from 55° F. (by the following ratio [1050/1155]×25 delta T=18° F. delta T. Furthermore, employing lower 18° F. delta T with the same AHU cfm supply air rate would still provide an equivalent and sensible cooling effect with a [75° F.−18° F. delta T]=57° F. AHU supply air temperature to meet required temperature of 75° F. at a 1155 cfm return air rate to the AHU.

Since the retrofitted or new sensible cooling only secondary AHU discharge temperature can be automatically reset by our PLC to 57° F. (from 55° F.), the net energy saved by the AHU amounts to 4% of input to a chiller (or equivalent refrigeration compressor energy saving) from the resulting 2° F. rise in AHU supply air temperature. It is believed that a benefit of 2% per 1° F. rise in AHU cooling coil discharge will be obtained.

Finally, assuming a 0.7 KW/ton input chiller (or equivalent refrigeration compressor) energy savings, based on the stated requirement, 58,590/12000 Btu/ton or 4.88 tons; 4%×4.88×12,000=2342 Btu/hr savings; therefore, total net energy saved=10,042 Btu/hr reduction in energy use achieved through the use of the dynamic OSA management system: [58,950−48,908] or 10,042 Btu/hr+[350−245] (or 595 cfm OSA at 0.0875 Btu/#BDA delta h) or 2342 Btu/hr]}/58,590=21% net energy saving.

The dew point detector 94 detects the dew point of the conditioned space or other system at the utilization location 24 and delivers this information to the PLC 72. The PLC also computes the dew point of the outside air in response to signals from the dry bulb and wet bulb thermometers 80 and 82.

By comparing the dew point of the outside air with the desired dew point, the PLC determines how much, if any, moisture is to be removed by the first evaporator DX2, and determines the necessary speed of the desiccant wheel, and controls the desiccant wheel motor 88 accordingly.

The performance characteristics of the desiccant wheel at different wheel speeds, air flow rates and moisture contents and moisture balance characteristics, are stored in the memory of the PLC, or in software used to operate the controller, so that the operational parameters to produce air with a desired output are developed and used to control the air temperature produced by DX2, wheel speed, etc.

Preferably, the exit temperature from the first condenser, DX2 is not substantially lower than that necessary to limit the desiccant wheel operation to a level which requires either no or minimum of supplemental heating of regeneration air.

By minimizing the cooling required from the first evaporator DX2, its size and operating cost can be minimized, and the temperature of the air leaving the desiccant wheel will be relatively high, creating a consistent need only for sensible cooling in the last stage of processing. No moisture removal is required from the evaporator DX1.

If the system shown in FIG. 2 is used, the second expansion valve 58 is adjusted by the PLC to give a variable but desired amount of cooling, so as to produce the desired temperature at the utilization location 24.

Based on the amount of moisture being removed by the desiccant wheel, based on the dew point outputs and the speed of the wheel, the supplemental heater 42 is turned on when the heat provided by the desuperheater 40 is insufficient to regenerate the desiccant material.

In either case, essentially no dehumidification is provided by the second evaporator, DX1; it, instead, provides only sensible cooling. Thus, the control of temperature and humidity are basically independent of one another, and the need for heating the air delivered is avoided.

By delivering air in a “space neutral” condition, the air can be introduced into different heating/cooling/dehumidification zones at a temperature and dewpoint close to that needed in that zone. Thus, the air is best adapted to the needs in each of many different zones.

FURTHER EMBODIMENTS

Further embodiments of the invention are shown in FIGS. 4-9 of the drawings. These drawings show a sustainable and energy efficient outside air desiccant—assisted pre-conditioner or conditioner system particularly suitable for use in microclimates with high prevailing seasonal humidities in excess of 100 grains/lb bone dry air (gr/#BDA).

The system, when used in the cooling mode as a preconditioner, also can serve to provide dehumidified outside air to new or existing package rooftop or interior air conditioner (AC) or (HVAC) heat pump unit. It may be configured as an integral component of either type of unit or installed to supply dehumidified air to one or more existing air conditioner or heat pump units either by direct ducted connection or connection in parallel when ducted to a common return plenum or supplied directly into conditioned space at a space neutral or variable discharge temperature. When interconnected with a new or existing heat pump it can be also be employed to maintain a minimum % relative humidity (% RH) during HVAC “heating” mode, for example.

One can practice the invention by using the system and method of the invention in any of several different configurations. For example, the system can be operated as a stand-alone rooftop unit, or as a package fully integrated and self contained (or split system configured) rooftop air conditioner or heat pump system. Alternatively, the system can be used as the preconditioner component of a system where separate secondary means is provided for sensible cooling to deal with both seasonal and internally generated cooling requirements. When required, in such a system, supplemental space heating or cooling can be provided when operating in a re-circulating OSA, return and exhaust constant volume or VAV air distribution system.

EXAMPLE

Assume that one wishes to condition 4,000 scfm of outside air in a 100% OSA dedicated outside air system as shown in FIG. 4. It is used in the cooling mode in a climate requiring much dehumidification and little or no heating year-round.

The components of the system shown in FIG. 4 are the same as those shown in the upper halves of FIGS. 1 and 2, except that heat-pipes HP are used as a heat exchanger instead of the sensible wheel 30, and the supplementary heater 42 is not shown in FIG. 4.

In FIG. 4, the air streams of principal interest are individually numbered 1 thru 11. The system parameters in each of those air streams for a 4,000 SCFM flow is shown in the table below:

#	SCFM	DB (° F.)	W (Gr/lb)	H (Btu/lb)
1	4,000	95	145	45.6
2	4,000	68	100	31.9
3	4,000	111	47	34.1
4	4,000	86	47	28.0
5	4,000	73	47	24.2
6	4,000	75	47	24.7
7	3,400	75	65	28.2
8	3,400	100	65	34.3
9	3,400	147	65	45.7
10	3,400
11	3,400
Note:
desiccant wheel cassette of 1220 mm dia. X 200 mm depth, rotating at 40 RPH

• • Total process (1-5) Q_Total=4000×4.5×(45.6−24.2)=385,200 Btu/hr=32.1 RT
  • DX2 (1-2) Q₂=4000×4.5×(45.6-31.9)=246,600 Btu/hr=20.6 RT
  • DX1 (4-5) Q₁=4000×4.5×(28.0−24.2)=68,400 Btu/hr=5.7 RT
  • DX2+DX1 Q_REFRIG.=Q₂+Q₁=20.6+5.7=26.3 RT

In the foregoing table;

“SCFM” is the outside air flow rate, in standard cubic feet per minute;

“DB” is the dry bulb temperature of the air, in degrees F.;

“W” is the humidity ratio, in grains per lb;

“H” is the enthalpy, in Btu/lb; and

“Q” is the cooling capacity required, in tons.

As it can be seen, outside air enters at 95° F. dry bulb and 145 grab humidity corresponding to an enthalpy of 45.6 Btu/lb, a condition (or state) also represented in the psychometric diagram of FIG. 5 by the same designated air flow arrow numbers.

In the table above, Qtotal corresponds to the mechanical refrigeration needed in tons (RT) to condition the 4000 scfm of OSA to obtain the desired air stream 5 conditions of 73F and 43 gr/lb. or 32.1 RT. Therefore to maintain conditioned space parameters of 75° F. and 50% RH, the preconditioned OSA would need to enter the space independently or via another means at air location 6, or at a space neutral DB temperature (75° F.) after adjustment for fan reheat and any ductwork distribution, sensible heat gain (assumed for our purposes to be 2° F.) and have enough latent heat capacity at 47 gr/lb to maintain 50% RH in the conditioned space. This assumes that any sensible heat gain due to people, lights, equipment and associated peak building envelope sensible transmission heat gains would be removed by independent cooling means dedicated to removing only sensible heat gains.

The comparable mechanical refrigeration capacity (shown above as Qtotal) is believed to be reduced by 5.7 RT or a total Q1+Q2 or approximately 18% by use of the invention.

FIG. 5 is a psychometric chart of air stream flow state points showing the results of the above. The large numbers correspond to the air stream numbers in FIG. 4.

It should be understood that although not shown in FIG. 4, a supplemental heater 42 as shown in FIG. 1 is provided in the regeneration air stream immediately preceding the desiccant wheel DH whenever the heat available from the desuperheater DSH falls below 147° F., at the conditions given above for this example, so as to maintain the humidity at 47 gr/lb, or whenever the compressor cycles off, for example, when both the thermostat 84 and dewpoint detector 94 indicate a need for dehumidification.

Additionally, although not shown in the drawings, if it is advantageous to do so, other cooling equipment such as indirect evaporative coolers can be used to cool the outside air before dehumidification. Also, enthalpy wheels or other heat exchangers can be used to exchange heat between the outside air and the exhaust air before the outside air is de-humidified.

FIG. 6 is a refrigeration system schematic diagram for the system of FIG. 4. The components shown in FIG. 6 are the same as in FIG. 1, except that the heat exchanger HTHX and the two valves “CV” have been omitted from FIG. 1 as being unnecessary. The heat exchanger “LTHX” in FIG. 6 is the same as the subcooler 52 shown in FIG. 1. Liquid flow paths are shown as solid lines and gas or gas-liquid flow paths are shown as dashed lines. The air flow at location numbers 1-2, 4-5 and 8-9 are the same as in FIG. 4. The valves “DV” are air-coil discharge sensor-activated three-way temperature control valves for air cooling coils DX1 and DX2.

The lower case letters in FIG. 6 indicate refrigerant temperature-pressure-state conditions as identified on the schematic Temperature-Entropy or TS diagram of FIG. 7.

FIG. 7, as noted above, is a TS diagram describing the refrigeration cooling cycle of the FIG. 6 system, and using the same lower case letters identifying refrigerant state conditions or state points at the corresponding points in FIG. 6. In FIG. 7, it should be noted that the process, between points e and f, is assumed to be isentropic expansion (i.e., with constant entropy).

Heat-Pump Embodiments

FIGS. 8 and 9 show operation of the system shown in FIGS. 4 and 6, and those of FIGS. 1 and 2 when operating in heat-pump mode.

This mode of operation is preferred for use in northerly temperate climates such as that in Boston, New York, Chicago, USA, or wherever operation in both heating and cooling (and free-cooling) modes is needed or desired.

Cooling Mode

FIG. 8 shows the system of FIG. 6 modified for operation as a heat-pump in the cooling mode. FIG. 8 is the same as FIG. 6 with the exception of the insertion of a 4-way solenoid valve (shown at lower left hand corner of FIG. 8) used to switch the heat pump from the cooling and de-humidification mode to the heating and humidification mode by redirecting refrigerant flows as shown FIG. 9. This reverses the functions of the condenser and evaporators DX1 and DX2 and by-passes flow to the desuperheating coil DSH as well, when used in the heating mode.

The valves CV shown in FIGS. 8 and 9 are optional bypass valves. They can be used for re-directing the flow of refrigerant to enable the use of a fixed-output compressor to give variable output, or for the purposes of enabling alternative interconnections.

Heating Mode

FIG. 9 shows the system of FIG. 8 in the heating mode. Refrigerant flow is reversed from that shown in FIG. 8 by the 4-way solenoid valve so that DX1 now serves as a low temperature (LT) condenser and DX2 now serves as a high temperature (HT) condenser and the former condenser now becomes the evaporator. The evaporator extracts heat from ambient air and/or exhaust air to provide heat for humidification and tempering of the OSA to deliver air at approximately 72° F., partially humidified. Thus, there is provided improved comfort and static electricity-free air but at a humidity below that needed to result in moisture condensation on exterior glass.

For example, humidification occurs by enabling the DH wheel to operate with the desuperheater DSH by-passed thus, in effect, transferring moisture from the return air to the OSA air. The air then is heated by LT condenser DX1 to the desired delivery temperature.

Operation of the motor-driven DH wheel will be cycled by humidity detected either in the OAS supply duct or the dew point controller 94 located in conditioned space to maintain safe humidification levels and avoid condensation on exterior envelope wall or glazing.

FIG. 10 is a schematic diagram of the heating-mode heat-pump system 100, of FIG. 9, with the mechanical components added, as in FIGS. 1 and 2.

The 4-way solenoid valve 101 has two different flow paths 102 and 104. When the valve 101 is in the state shown in FIG. 10, refrigerant flows through path 102 from LTHK (Low Temperature Heat Exchanger) 52, which is the same as the subcooler shown in FIGS. 1 and 2, to DX2, which now has been connected to become a condenser to be a heater instead of a cooler.

The HTHX (High Temperature Heat Exchanger 106) is interconnected with the LTHX and other components as shown in FIG. 9.

The desuperheater 40 is disconnected and inactive, as is the supplemental heater 42. The sub-system 108 is for heating, and is not a refrigeration sub-system as is the system 21 in FIGS. 1 and 2.

The valves CV shown in FIGS. 8 and 9 have been omitted from FIGS. 10-12.

FIG. 11 shows a system 110 that is the same as that in FIG. 10, except that the desuperheater is connected by lines 112 and 114 to receive refrigerant and thus become a second evaporator unit to supplement the unit 50 and increase the output of the system.

FIG. 12 shows the heat-pump system 120 used in the cooling mode. The system 120 is the same as shown in FIG. 8, except that the mechanical components have been added, as in FIGS. 1 and 2.

The valve 101 now has changed state to direct the flow through path 104 to create the cooling mode of the heat-pump operation. The desuperheater 40 now is connected to heat the regeneration air, as in FIGS. 1 and 2. The units DX1 and DX2, and the condenser 50 have again reversed the functions they have in FIGS. 9-11 to be used in cooling and dehumidification.

The introduction of 10% to 15% return air at the outside air inlet is eliminated.

There are numerous HVAC air distribution options for removing conditioned space sensible (only) heat and provide heating as discussed above where required.

Mention was made above of numerous means of incorporating the active 100% OSA preconditioner as a component of one of the following 8 air distribution types or as integral part of any of the air distribution system where both indoor and outdoor latent humidity requirements for conditioned building spaces are controlled independently, thereby permitting sensible heat gains to be removed by the following supplemental sensible cooling only downstream companion separate systems; namely:

• • (1) fan coil air distribution units,
  • (2) mini-split ductless air conditioning units,
  • (3) air handlers operating at higher temperature to achieve a more energy efficient DX evaporator coil operation
  • (4) chilled water cooled coil air handlers with either:
    • (4a) a fixed OSA ventilation rate or
    • (4b) a constant volume air distribution system, with OSA, return and exhaust air streams, or
    • (4c) variable air volume (VAV) air distribution system, with OSA, return and exhaust air streams
  • (5) 100% OSA DOAS type distribution systems also incorporating chilled water
  • (6) mini-split ductless fan powered terminals
  • (7) series fan powered terminals for secondary space sensible (only) cooling needs all of which can be controlled from zoned thermostats while also resulting in improved indoor IAQ at all part load space occupant cooling
  • (8) parallel fan powered terminals for secondary space sensible (only) cooling needs all of which can be controlled from zoned thermostats while also resulting in improved indoor IAQ at all part load space occupant cooling and heating requirements
  • (9) heat pump unit heated and humidified exterior space heating conditions all of which have been shown in FIGS. 8 and 9.

In operation, in the cooling mode, the first cooling coils DX₂ (FIG. 4) receiving OSA are deliberately set to cool incoming air to a temperature, preferably above the conditioned space desired temperature, so that little or no dehumidification takes place.

Claims

1. An outside air supply management system for a building, said system comprising

(a) first cooling apparatus for cooling incoming outside air to a variable first temperature, at which a quantity of moisture is removed from said outside air, said quantity varying from zero to a substantial quantity,

(b) a variable capacity desiccant wheel controllable for removing from said air substantially all of the remainder of the moisture designed to be removed from the air received from said first refrigeration unit and issuing dehumidified air at a desired dewpoint, and a temperature above a desired temperature to be maintained at a utilization location,

(c) a second cooling apparatus for cooling air from said desiccant wheel to a second temperature desired to be maintained at a utilization location,

(d) at least one air mover for moving said outside air through said first cooling apparatus, said desiccant wheel, and said second cooling apparatus to said utilization location, and

(e) a programmed controller for controlling the operation of said cooling apparatus and said desiccant wheel to achieve the temperatures and dewpoint specified above and thereby control the humidity and temperature of the air delivered to said utilization location substantially independently of one another.

2. A system as in claim 1 in which each of said first and second cooling apparatus comprises a refrigeration evaporator, and a single compressor supplying both of said evaporators with refrigerant, and at least one expansion valve controlled by said controller to govern the cooling provided by said evaporators, and a flow re-directing valve operable to convert said system between a cooling system and a heating system in which said evaporators are converted into condensers used for heating.

3. A system as in claim 1 including a desiccant regeneration sub-system comprising an air mover for moving return air from said utilization location to heating apparatus, said heating apparatus, including a desuperheater and then to said desiccant wheel to regenerate the desiccant therein, said programmed controller being programmed to set said first temperature at a level sufficient to assure regeneration of the desiccant in said desiccant wheel.

4. A system as in claim 1 in which said cooling apparatus includes at least two evaporators are connected together in a refrigerant flow circuit selected from the group consisting of a parallel connection with a single expansion valve, and a tandem connection with two separate expansion valves, one for each of said evaporators.

5. A system as in claim 1 including a conduit for introducing a portion of said return air into the incoming outside air.

6. A system as in claim 1 including a heat exchanger forming part of said second cooling apparatus for absorbing heat from the air leaving said desiccant wheel and delivering heat to said return air to heat it before reaching said heating apparatus and said desiccant wheel.

7. A system as in claim 1 including equipment for sensing the dewpoint of air in said utilization location, a further dew point detecting equipment for said outside air, said dew point equipment being connected to provide inputs to said controller, the performance characteristics of said desiccant wheel being stored in said controller for use in controlling the speed of said desiccant wheel, and compressor control equipment for changing the output of a compressor in said system to modulate cooling in said system.

8. A system as in claim 1 including a thermostat and a CO2 sensor in said utilization location and delivering signals to said controller, and a desiccant wheel rotator motor connected to be controlled by said controller in accordance with the desiccant moisture reduction needed, a sensible cooling air handling unit receiving conditioned air outside from said outside air supply system, said controller being programmed to increase the internal air temperature design point in response to reductions in outside air volume requirements.

9. A system as in claim 1 in which each of said first and second cooling apparatus includes an evaporator, and including a variable compressor for compressing refrigerant gas and supplying it to both of said evaporators, a desuperheater, at least one condenser, a subcooler, and at least one controllable valve for selectively operating said cooling apparatus to produce variable amounts of cooling as needed.

10. A system as in claim 9 in which the refrigerant used in the refrigeration system in which said evaporators are used is R22 or equivalent.

11. A system as in claim 3 in which said system includes an evaporator for each of said cooling apparatus, a condenser and a compressor, and is part of a heat-pump system with at least one valve for selectively reversing the flow of refrigerant through said evaporators and condenser and reverse the functions of said condenser and said evaporators and convert said system from a cooling mode to a heating mode.

12. A system as in claim 11 in which said desiccant wheel is operated to transfer moisture from said return air to said incoming outside air for humidification, with said evaporators converted to heaters, and said system includes a desuperheater which is connected for reverse operation as an evaporator.

13. A method of conditioning outside air to cool and dehumidify said outside air and deliver the cooled and dehumidified air to an internal utilization location in a building, said method comprising the steps of:

(a) moving outside air into contact with a first variable cooling device and controlling the temperature produced in said outside air to a first temperature to remove from zero to a substantial portion of the moisture to be removed from the air,

(b) moving air from said first cooling device to a desiccant device and controlling said desiccant device to remove from substantially all of the remaining moisture desired to be removed therefrom,

(c) removing heat from the air leaving said desiccant wheel by means of a second cooling device to a temperature desired for said utilization location, and

(d) moving the air leaving said second evaporator to said utilization location.

14. A method as in claim 13 including the steps of:

(e) moving exhaust air from said utilization location to a heat exchanger to absorb heat removed from air leaving said desiccant device,

(f) moving exhaust air from said heat exchanger to a desuperheater and heating said air therein,

(g) using the heated exhaust air from said desuperheater to regenerate the desiccant material in said desiccant device, and

(h) controlling said first temperature so as to eliminate or minimize the need to supplementally heat said exhaust air.

15. A method as in claim 13 including using a programmed electronic controller to control the amount of cooling provided in such of said cooling steps and the amount of moisture removed from the outside air by said desiccant device to deliver outside air to a utilization location at a desired temperature and humidity with each quantity being controlled substantially independently of the other.

16. A method as in claim 13 including using a programmed electronic controller to control the amount of cooling provided in such of said cooling steps and the amount of moisture removed from the outside air by said desiccant device to deliver outside air to a utilization location at a desired temperature and humidity substantially without reheating said air, and utilizing said controller to compute the level of operation of each of said steps to produce said outside air at the desired temperature.

17. A method as in claim 13 including sensing CO2 levels in the utilization location and delivering corresponding signals to control the steps of moving air to provide the amount of ventilation needed as indicated by said CO2 levels and increasing the required temperature level of a conditioned space as a function of the reduction in outside air volume.

18. A method as in claim 13 including programming a programmable controller to provide set levels of parameters based on the time of year and time of day in the location of the building of said utilization location

19. An air conditioning system particularly well adapted to condition extremely high-humidity air, said system comprising:

(a) a first DX coil for cooling incoming outside air to a first temperature,

(b) a desiccant wheel for dehumidifying air received from said wheel,

(c) a heat exchanger selected from the group consisting of a heat pipe structure and a sensible heat wheel receiving air from said desiccant wheel and transferring said heat to another location spaced from the path of said air leaving said desiccant wheel,

(d) a second DX coil for cooling air received from said heat exchanger and delivering the cooled air to a utilization location.

(e) a fan for moving air from said utilization location through said other location to a desuperheater and then to the regenerative side of said desiccant wheel to regenerate the desiccant therein, and

(f) a device for controlling the amount of cooling provided by said first DX coil to reduce said first temperature to near the saturation level of said outside air.

20. A system as in claim 19 including a condenser, a single compressor supplying both of said DX coils, at least one expansion valve, and a valve for selectively reversing the flow of refrigerant through said DX coils and said condenser to provide for heating operation of said system, with said desiccant wheel selectively transferring moisture from return air to said outside air for humidification.