US20050235666A1

US20050235666A1 - Integrated dehumidification system

Info

Publication number: US20050235666A1
Application number: US11/115,188
Authority: US
Inventors: David Springer; Marc Hoeschele; Mark Berman; James Phillips
Original assignee: Davis Energy Group Inc
Current assignee: Davis Energy Group Inc
Priority date: 2004-04-27
Filing date: 2005-04-27
Publication date: 2005-10-27

Abstract

A dynamic system controls indoor relative humidity and temperature to achieve specified conditions by applying multiple stages of dehumidification. In addition to an optional stage that increases dehumidification by reducing the speed of the indoor blower, the system uses a reheat coil and multiple valves that allow the reheat coil to function as either a subcooling coil or a partial condenser. Thus the system can maintain specified indoor temperature and humidity conditions even at times when no heating or cooling is needed. The system may have an outdoor condensing unit including a compressor and a condenser operably connected via refrigerant lines to an indoor unit to form a “split system” refrigerant loop.

Description

CROSS-REFERENCE TO RELATED APPLICATION

This application claims the benefit of U.S. Provisional Application No. 60/565,532, filed on Apr. 27, 2004.

BACKGROUND

The subject matter of this disclosure relates to providing building cooling, dehumidification, and fresh air ventilation through a range of outdoor and indoor conditions.
New U.S. homes that are built in compliance with ASHRAE Standard 90.2, Energy Star, and other energy efficiency programs have lower cooling loads than in the past, and because they are of tighter construction, they frequently require mechanical ventilation as prescribed by ASHRAE Standard 62.2. In humid climates the ventilation air often requires more dehumidification than can typically be provided by air conditioners, because typical air conditioners in energy-efficient homes have short run times during many cooling load hours. Short run times typically limit latent cooling capacity. Failure to control excessive indoor humidity has contributed to problems with indoor mold. This issue has become increasingly expensive for homeowners and builders, as mold-related property damage and class action lawsuits have risen steadily.
Vapor compression cooling systems (air conditioners) that are in use in most homes and other buildings provide a mix of sensible cooling (lowering the air temperature) and latent cooling (removing moisture). Typically, the sensible heat ratio (“SHR”, the sensible cooling capacity divided by the total capacity) for most residential cooling systems ranges from 0.7 to 0.8. In humid conditions this SHR is often too high to maintain temperature and relative humidity in the ideal ranges of 74°-78° F., and 40-60%, respectively. Some vapor compression cooling systems lower the airflow rate through the evaporator coil to reduce the SHR under humid conditions, but re-evaporation of condensate retained on the coil at system shutdown still limits the SHR, particularly when systems cycle frequently, as they do under low load conditions. Such residential cooling systems are “split systems”, with an outdoor condensing unit that includes the compressor, condensing coil, and condenser fan, and a separate indoor unit that includes an evaporator coil, expansion device, and system blower. Two refrigerant lines join the outdoor and indoor components.
Furthermore, a stand-alone dehumidifier is frequently used in humid climates to control indoor humidity. Because heat from the condenser is added to indoor air, the dehumidifier often increases the sensible cooling load, the air conditioner run time, and the amount of energy consumption. A preferred approach to dehumidification in the cooling season is to dehumidify indoor air by rejecting condenser heat to outside air instead of to the indoor space.
In the prior art, various strategies have been proposed to control both temperature and humidity. For example, U.S. Pat. No. 6,170,271 B1 shows a concept with two separate refrigerant loops; a first loop with the evaporator in the supply air stream and the condenser outdoors, for sensibly cooling the air stream, and a second “latent cooling” loop with the evaporator just downstream of the first evaporator, and with the condenser downstream of the second evaporator. This approach is similar to combining an air conditioner and a dehumidifier, but with the added benefits of requiring only one indoor blower and cabinet, and a smaller second evaporator can be used because the air has been pre-cooled in the first evaporator. However, all heat from the second loop is added to the supply air, with associated energy penalties. In the embodiment, having the dehumidifier condenser located outside the supply air stream, the system is still penalized by the cost of requiring dual compressors, additional refrigerant piping, and condensers. Various other design configurations appear in the patent literature and are aimed at more precisely controlling both sensible and latent loads.
Another strategy having dual refrigerant loops is shown in U.S. Pat. No. 6,705,093 B1 and uses two condensing units that share an evaporator coil whose tubing pattern maintains separation of the two loops. One of the two loops has a sub-cooling coil. This approach adds substantial cost to a conventional system with a single refrigerant loop. Another approach to increasing latent cooling is shown in U.S. Pat. No. 6,427,454 B1. This design selectively causes a portion of the return air to bypass the evaporator coil, which lowers the coil temperature and increases moisture condensation on the coil. However, this approach is unlikely to succeed in the market, as it is comparable to lowering the blower speed, but with higher initial costs and without the energy savings associated with reducing blower speed.
U.S. Pat. No. 6,123,147 shows a retrofit system that adds a hot water reheat coil connected to a residential water heater located downstream of the evaporator. Like other “reheat” designs, this approach decreases the SHR by making the cooling system run longer. However, the economics of such a system will be poor because gas water heating is substituted for waste heat already available from the condensing side of the refrigerant system. Thus, this approach is like driving a vehicle using the accelerator and brake simultaneously. Other strategies, such as that disclosed in U.S. Pat. No. 5,791,153, apply desiccant-based enthalpy wheels to increase latent cooling. These designs require added components to recharge the desiccant and therefore may not be cost-effective.
Of the major product lines in the U.S. marketplace, only the Carrier® Infinity™ series and the Lennox™ SignatureStat™ controller claim features that control both temperature and humidity. However, both products can only control humidity by varying fan and compressor speed. There are no added components designed to respond to conditions with high humidity and low cooling loads. Thus, these systems cannot maintain a specified temperature/humidity set through a wide range of conditions.
In the “packaged” air conditioning market with products usually applied to non-residential buildings, Lennox™ markets a patented Humiditrol® line that includes refrigerant control valves and a “hot gas” reheat coil for more precise humidity control. Carrier® markets the MoistureMiser™ that uses a “sub-cooling” coil for the same purpose. In both cases the strategy is to add some heat from the condenser side of the refrigerant system back into the supply air stream (downstream of the evaporator) to reduce the net cooling rate. Such systems must run longer to satisfy the cooling load, and the longer run time removes more moisture at the evaporator. Adding more length to the coil on the condenser side also reduces the liquid refrigerant temperature into the evaporator, which increases evaporator capacity and therefore drops the evaporator temperature, increasing the rate of moisture removal. Lennox™ claims superior dehumidification performance because the higher heat output of the “hot gas” approach causes longer cooling cycles, thus removing more moisture compared to the sub-cooling approach.
These non-residential products use a “single-package” configuration, and no “split system” units currently include the “reheat coil” features described above. In fact, the Lennox™ hot gas approach is only workable in a single package device, as the system would require an extra pair of refrigerant lines to be applied in a “split system” configuration because refrigerant must flow first to the indoor reheat coil, then back to the condenser, then to the indoor expansion device. The Carrier® sub-cooling approach would not require an extra line set in a split system configuration because the refrigerant flows directly from the sub-cooling coil to the expansion device. However, the approach only provides two stages of dehumidification, and therefore cannot sufficiently control humidity when sensible loads are very low and latent loads are high.
Although the vapor compression systems disclosed above, and others, use hot gas and sub-cooling reheat coils to reduce the SHR in single-package units, no known systems dynamically combine features that, by applying multiple dehumidification stages in a split system configuration, can maintain desired temperature and humidity conditions even in the absence of cooling loads, through the full range of climatic conditions in the U.S. and elsewhere.

SUMMARY

The most desirable indoor comfort system for humid climates would use minimal added components to a conventional split air conditioning system, but would have the capability of dehumidifying even in a “neutral” condition wherein the building needs neither heating nor cooling. Using a single refrigerant loop with a supplemental reheat coil to achieve this condition would require that the evaporator and the reheat coil have equal and opposite heat transfers to the air stream. The outdoor coil is then rejecting a heat quantity equal to the compressor input energy. A low indoor airflow rate is desirable to maximize latent cooling, using care not to freeze the evaporator coil.
In various exemplary embodiments, the systems and methods of this invention provide automatic, dynamic control of indoor relative humidity and temperature to achieve specified conditions by applying multiple stages of dehumidification. Various aspects of the exemplary embodiment include the capability to remove moisture from outside ventilation air supplied to maintain indoor air quality at times when no heating or cooling is needed. Still another aspect is the ability to maintain a specified indoor relative humidity through a wide range of climates and seasonal conditions. For economic viability, such systems should readily integrate with conventional heating and cooling components, applying the fan, coil, and condensing unit to both sensible cooling and dehumidification functions.
In various exemplary embodiments, the systems and methods of this invention surpasses the efficiency of air conditioners combined with stand-alone dehumidifiers, by rejecting condenser heat developed in the dehumidification process to outdoors instead of indoors.
In various exemplary embodiments, the systems and methods of this invention efficiently and effectively dehumidifies outside ventilation air supplied to buildings for the purpose of maintaining indoor air quality.
In various exemplary embodiments, the systems and methods of this invention combines indoor cooling and dehumidification components into a single unit to facilitate installation and reduce cost.
These and other objects and advantages will be apparent to those skilled in the art in light of the following disclosure, claims and accompanying drawings.

BRIEF DESCRIPTION OF THE DRAWINGS

The invention will be described in detail with reference to the following drawings in which like reference numerals refer to like elements and wherein:
FIG. 1 is a schematic diagram of the refrigeration and control components of the invention showing alternate refrigeration flow paths for the various dehumidification stages;
FIG. 2 is a schematic diagram of the refrigeration and control components of the invention showing alternate refrigeration flow paths for the various dehumidification stages; and
FIG. 3 is a schematic diagram of the refrigeration and control components of the invention showing alternate refrigeration flow paths for the various dehumidification stages.

DETAILED DESCRIPTION OF EMBODIMENTS

An exemplary embodiment of the systems and methods described in this disclosure comprises a set of vapor compression cooling components that can respond to a wide range of sensible and latent cooling loads, and control components with appropriate logic for automatically maintaining indoor temperature and relative humidity within close tolerances. The embodiment can condition either re-circulated indoor air, outside ventilation air supplied to buildings to maintain indoor air quality, or a mixture of the two. Exemplary components of such a system include a compressor, a condensing coil, a condenser fan, an indoor blower, an evaporator coil, a reheat coil, a refrigerant receiver, a thermostatic expansion valve, solenoid valves for switching refrigerant flow, a check valve, “pressure-differential check valves” (PDCV's), temperature and humidity sensors, and controls for selecting an operating mode based on sensed conditions.
With reference to FIG. 1, an integrated dehumidification system 100 comprises an outdoor-condensing unit 1, an indoor unit 40, refrigerant lines 7 and 13 that connect the condensing unit 1 and the indoor unit 40, and a control system 30.
The condensing unit 1 includes a compressor 2, condensers 3, a cabinet 4, and a condenser fan 5 driven by a condenser fan motor 6. Major components of the indoor unit 40 includes an evaporator coil 12, a reheat coil 8, a blower 21 driven by motor 22, automatic valves 14 and 15, and an enclosing cabinet 41. The indoor unit also includes PDCV's and refrigerant lines as will be discussed with respect to the specific dehumidification stages. The control system 30 includes a thermostat and logic board 33, switching/relay boards 34 in the indoor unit and 35 in the outdoor condensing unit 1, and indoor air sensors for temperature 31 and humidity 32, and an optional coil freeze sensor 37.
In an exemplary embodiment, the integrated dehumidification system 100 includes four dehumidification modes. In an exemplary embodiment, a “Stage 1 dehumidification” mode has the lowest latent cooling capability and the highest SHR and may use a refrigerant flow schematic similar to that for a conventional split air conditioning system. Low pressure refrigerant vapor is compressed to a superheated, high pressure vapor state in the compressor 2 of the outdoor condensing unit 1. The vapor then passes through the condenser coils 3 where the vapor condenses to a liquid state, giving up heat, before leaving the outdoor condensing unit 1 through the refrigerant line 7. During this process, the condenser fan 5, driven by the fan motor 6, induces outdoor airflow across the condenser coils 3 to discharge heat to outdoor air. Although FIG. 1 shows two condenser coils 3 in parallel, one “wrap-around” coil may be used as well.
After the liquid refrigerant enters the indoor unit 40 through the refrigerant line 7, the liquid refrigerant passes through an open automatic control valve 14. In the exemplary embodiment, there are multiple parallel paths through lines 9, 17, 19, and 42, toward the evaporator coil 12. But all of these paths are blocked by either a check valve 36 or PDCV's 16 a, 16 b that have pressure drop settings higher than the downstream pressure drops between the entering refrigerant line 7 and an expansion device 11. After passing through the automatic control valve 14, the refrigerant flow proceeds through the line segment 23 into the liquid receiver 10, then through another open automatic control valve 15 via the line segment 20, and on through the expansion device 11. The expansion 11 restricts refrigerant flow and causes the high pressure liquid to begin a change of state from a liquid to a low pressure gas. From the expansion device 11, the refrigerant enters the evaporator coil 12 where the change of state is completed. As the refrigerant evaporates at the evaporator coil 12, the refrigerant absorbs heat from the air stream 26 driven through the air path 18 across the evaporator coil 12 by the indoor blower 21 powered by the blower motor 22. The heat absorbed by the refrigerant results in cooling of the air stream 26. If the surfaces of the evaporator coil 12 are cooler than the dew point temperature of the air stream 26, moisture will condense on the coil 12 and drip into a drain pan 27 from which it can be drained through condensate drain 28. From the evaporator 12 the low pressure refrigerant vapor returns through the refrigerant line 13 to the compressor 2 of the outdoor condensing unit 1.
With continuing reference to FIG. 1, “Stage 2 dehumidification” mode of the exemplary embodiment uses a “reduced air flow” strategy. In Stage 2, the speed of the blower motor 22 is reduced, thereby reducing the flow rate of the air stream 26. The reduced speed of the blower motor 22 increases air stream residence time and causes a reduction of the evaporating temperature in the evaporator coil 12, thereby increasing dehumidification. The control system 30 is programmed with staged thresholds for indoor relative humidity. For example, when a first user-selected threshold is exceeded, the control system 30 will shift the operating speed of the blower motor 22 from a normal speed to a programmed lower speed. If indoor humidity later drops slightly below the first user-selected threshold, the control system 30 returns the operating speed of the motor 22 to the normal speed.
FIG. 2 shows the refrigerant flow in the indoor condensing unit 40 when a second humidity threshold is exceeded. In this “Stage 3 dehumidification” mode the automatic control valve 14 remains open, and the refrigerant flow passes through the receiver 10 as in Stages 1 and 2. In Stage 3 the refrigerant then flows through the line 42 toward the reheat coil 8, rather than through the line 20 toward the evaporator coil 12, because the automatic control valve 15 in the line 20 is now closed. A PDCV 16 a that requires approximately 5 psi of pressure to overcome its spring force is located between the refrigerant line 7 and the intersection of the lines 42 and 9 to prevent the refrigerant from flowing directly into the reheat coil 8 in the first three dehumidification stages. A check valve 36 in the line 19 prevents bypassing of the reheat coil from the line 23 above the receiver 10 to the line 17 toward the expansion device 11. From the exit of the reheat coil 8, all refrigerant flows through the line 17 and through PDCV 16 b to the expansion device 11 and the evaporator coil 12 before completing the circuit back to the compressor 2 of the outdoor condensing unit 1 through the refrigerant line 13. In this circuit, the liquid refrigerant from the condenser 3 (see FIG. 1) is sub-cooled in the reheat coil 8. This process increases dehumidification mostly by adding heat back into the air stream 26 downstream of the evaporator coil 12, which reduces the cooling delivery rate and causes the dehumidification system 100 to run longer to satisfy the cooling load. Longer operation with a constant surface temperature pattern for the evaporator coil 12 results in more moisture removal as long as part of the surface of the evaporator coil 12 is colder than the dew point temperature of the entering air stream 26. This circuit offers an additional dehumidification benefit by sub-cooling the liquid refrigerant below the condensing temperature to lower the evaporating temperature and thus increase the rate of moisture removal. The control system 30 (see FIG. 1), implements Stage 3 dehumidification by closing the automatic control valve 15 when a second user-selected threshold is exceeded.
If the humidity sensor 32 (see FIG. 1) indicates that a third user-selected threshold has been exceeded, the control system 30 will initiate a “Stage 4 dehumidification” mode operation as shown in FIG. 3. In Stage 4 mode, the automatic control valve 15 is opened and the automatic control valve 14 is closed so that the incoming refrigerant flow from the outdoor condensing unit 1 (see FIG. 1) is forced through the line 9 with a PDCV 16 a into the reheat coil 8. The flow then proceeds through a low pressure drop check valve 36 in line 19 before entering the receiver 10. The PDCV 16 b imposes a greater pressure drop in line 17 than the sum of the pressure drops in the lines 19, 20, the receiver 10, and the open valve 15. As a result, flow is forced through the receiver 10. From the receiver 10 the refrigerant flow proceeds through the open automatic control valve 15 in the line 20 and through the expansion device 11 before entering the evaporator coil 12. With the receiver 10 downstream of the reheat coil 8, the refrigerant can partially condense in the reheat coil 8 because the refrigerant will preferentially condense in the coldest available location. Because the reheat coil 8 is in the low temperature air stream 26 leaving the evaporator coil 12, the reheat coil 8 will be typically be cooler than the condensing coil 3 (see FIG. 1) located in outdoor air. As a result the refrigerant partially condenses in the reheat coil 8, delivering more reheat than was available in Stage 3 dehumidification mode.
In an exemplary embodiment, it is possible to operate in the Stage 4 dehumidification mode without either cooling or heating the supply air stream. In this “neutral” dehumidification case, sufficient condensing occurs in the reheat coil 8 to balance the cooling delivered at the evaporator coil 12, and the heat being discharged at the condensing unit 1 equals the equivalent heat input of the compressor 2 (see FIG. 1). In contrast, a conventional dehumidifier adds all heat, including the compressor input heat, to the space in which it is enclosed. Without applying controls at the condensing unit, there are two ways to accomplish the neutral dehumidification state. One is to combine a relatively small condensing coil 3 (FIG. 1) and a relatively large reheat coil 8. This sizing approach will probably compromise Stage 1 (normal cooling) operation. The other is to combine a normally-sized condenser coil 3 with a large reheat coil 8, and operate the blower motor 22 at a sufficiently low speed that air leaving the evaporator coil 12 is just above freezing temperature. (An evaporator coil surface temperature sensor 37 (see FIG. 1) should be used to increase the airflow rate when there is danger of freezing moisture on the evaporator coil 12.) This strategy maximizes dehumidification, and causes the lowest possible temperature entering the reheat coil 8, increasing the amount of refrigerant condensation that occurs in the reheat coil. However, this strategy may drop refrigerant pressures sufficiently to trip the low pressure cut-out-typically included in the condensing unit 1. Thus, care must be used in sizing the coils and compressor.
An aspect of such a dehumidification system is that no control interaction with the condensing unit is required. The system may also be coupled with any available condensing unit. However, applying added controls to the condensing unit components offers improved dehumidification control. Such condensing unit controls provide two additional strategies or stages of dehumidification that can further reduce the SHR without penalizing Stage 1 cooling operation. For example, a first strategy may be to couple the air handler 40 (see FIG. 1) with a two-speed condensing unit 1. A two-speed condensing unit typically includes a two-speed compressor 2 and a two-speed condenser fan motor 6. Control access to these components offers the opportunity for additional dehumidification benefits. For example, if a one-speed compressor moves into an unacceptably low pressure operating regime in the Stage 4 dehumidification mode, a solution may be to select a two-speed condensing unit and shift to low speed for the Stage 4 operation. Another potential benefit of control access to the condensing unit is the opportunity to reduce the speed of the condenser fan motor 6 to reduce condenser heat transfer in the Stage 4 dehumidification mode. In an extreme case, the condenser fan motor 6 can be disabled so that most of the condensation occurs in the reheat coil. The system will then operate nearly like a packaged dehumidifier, causing a net heat addition to the space.
With the multiple stage dehumidification strategies described here, it is possible to satisfy both temperature and humidity targets in indoor spaces through a full range of outdoor, indoor, and ventilation conditions. When control access to the condensing unit is available, the system can even dehumidify in the absence of cooling loads, or can deliver heat while dehumidifying if desired. In each stage of the dehumidification operation, the system can operate at maximum potential efficiency by rejecting the most heat possible to the outdoor environment while satisfying the indoor temperature and humidity targets.
Although the invention has been shown and described with respect to a preferred embodiment thereof, it should be understood by those skilled in the art that various changes and omissions in the form and detail thereof may be made therein without departing from the spirit and scope of the invention. For example, the system has been described assuming an air-cooled condenser, which is currently the most common condenser type. However, the multi-stage dehumidification strategies described here may as easily be applied with water-cooled condensers or storage-type condensers such as hydronic or direct refrigerant ground-loops.

Claims

1. A system for controlling indoor environmental conditions, comprising:

a vapor-compression refrigerant loop with a compressor, a condenser, a reheat coil, an expansion device, and an evaporator coil;

an air mover that drives an indoor air stream sequentially across the evaporator and a reheat coil;

a plurality of valves disposed within the vapor-compression loop; and

a controller operably connected to the vapor-compression loop, the air mover and the plurality of valves, wherein the condenser discharges heat to an outdoor environment, and the controller controls a plurality of dehumidification stages including a stage that causes refrigerant to condense in the reheat coil.

2. The system of claim 1, wherein the refrigerant loop includes a receiver for refrigerant volume control.

3. The system of claim 1, wherein one of the plurality of dehumidification stages minimizes dehumidification of indoor air by operating the air mover at a normal speed and directing a refrigerant in the refrigerant loop to bypass the reheat coil by operating at least one of the plurality of valves.

4. The system of claim 1, wherein one of the plurality of dehumidification stages increases dehumidification of indoor air by reducing a speed of the air mover to reduce a velocity of the indoor air stream across the evaporator and the reheat coil.

5. The system of claim 1, wherein at least one of the plurality of dehumidification stages protects the condenser from low refrigerant pressures by reducing a rate at which the condenser discharges heat to the outdoor environment.

6. The system of claim 1, wherein the condenser is cooled by a second air stream driven by at least one condenser fan and a flow rate of the second air stream is controlled by controlling the at least one condenser fan.

7. The system of claim 6, wherein the at least one condenser fan operates at a single speed and is cycled on and off to reduce a heat exchange rate of the second air stream.

8. The system of claim 1, wherein the controller is operably connected to an indoor temperature sensor, an indoor humidity sensor, at least one of the plurality of valves, and the controller includes control logic that controls the air mover, the condenser, and the at least one of the plurality of valves based on input from at least one of the indoor temperature sensor and the indoor humidity sensor.

9. The system of claim 2, wherein the refrigerant receiver is located between the condenser and the evaporator.

10. The system of claim 2, wherein the reheat coil is located in a first parallel path between the condenser and the expansion device, and the receiver is located in a second parallel path between the condenser and the expansion device.

11. The system of claim 10, wherein a first bypass path is provided between an outlet of the reheat coil and an inlet of the receiver.

12. The system of claim 11, wherein a second bypass path is provided between an inlet of the reheat coil and an outlet of the receiver.

13. The system of claim 12, wherein the plurality of valves includes first and second refrigerant valves placed in the second parallel path between the condenser and the expansion device and wherein the first refrigerant valve is placed between the condenser and the receiver, and the second refrigerant valve is placed between the receiver and the expansion device.

14. The system of claim 13, wherein the first bypass path joins the second parallel path between the condenser and the expansion device at a location between the first refrigerant valve and the receiver, and the plurality of valves includes a check valve located in the first bypass path to prevent refrigerant flow from the inlet of the receiver to the outlet of the reheat coil.

15. The system of claim 14, wherein the plurality of valves includes first and second pressure-reducing check valves located in the first parallel path between the condenser and the expansion device, and wherein the first pressure-reducing check valve is located between the condenser and the reheat coil, and the second pressure-reducing check valve is located between the reheat coil and the expansion device.

16. The system of claim 15, wherein the second bypass path joins the first parallel path between the first pressure-reducing check valve and the reheat coil, and the first bypass path joins the first parallel path between the reheat coil and the second pressure-reducing check valve.

17. A system for controlling indoor environmental conditions, comprising:

an outdoor condensing unit including a compressor and a condenser operably connected via refrigerant lines;

an indoor unit operably connected to the outdoor condensing unit to form a refrigerant loop, the indoor unit comprising:

an evaporator coil, a reheat coil, an expansion device, a refrigerant receiver for refrigerant volume control, an expansion device, and a plurality of valves, operably connected via refrigerant lines;

an air mover that drives a first air stream sequentially across the evaporator coil and the reheat coil; and

a controller operably connected to the outdoor condensing unit and the indoor condensing unit, wherein the condenser discharges heat to outdoor air, the reheat coil is located downstream of the evaporator coil in the first air stream, and the controller controls a plurality of dehumidification stages.

18. The system of claim 17, wherein the plurality of dehumidification stages includes a stage that causes refrigerant in the refrigerant loop to condense in the reheat coil.