I was invited to provide opinion regarding a newly installed air cooled water chiller system who's twin AHU's were apparently unable to properly manage the conditioned space's design humidity. It was also explained that the water chiller itself had been repeatedly tripping on low pressure safety (LP) since almost the first days of operation.

The cause turned out to be quite simple, at least more simple than usual as far as special troubleshooting requests go. It turned out that the chilled water piping had been reverse connected at the main return and supply headers, so we had the return water passing backwards, in a parallel flow configuration, through the chiller's shell and tube heat-exchanger, then through the chillers local strainer and on to the supply header. The resulting incorrect strainer position alone, situated after, rather than before, the heat-exchanger, might have been sufficient cause for frequent LP trips were there sufficient unstrained debris to foul the heat-exchanger.

It was also explained to me that as a result of the frequent water chiller LP trips, in attempt to solve the problem, the manufacturer had changed the two thermal expansion valve's (TEV's), one for each compressor. I noticed the replacement TEV's were a well known manufacturer's anti charge-migration type specially designed to isolate the bulb from the potentially colder power assembly with the use of a hydraulic-fluid charged capillary and power assembly. Charge migration from the TEV's bulb to the power assembly bellows can often occur when the TEV itself becomes colder than the TEV bulb, this causes lower than desired power assembly pressures which in turn results in excessive TEV throttling since the TEV is falsely seeing a reduced superheat.

If I'm permitted an attempt to elaborate on the details of this scenario, which may yet provide some entertainment.

The dynamics behind the LP trips.

In accordance with logical system component balance laws, subject to individual component operating characteristics, the minimum operationally established evaporator temperature difference (TD) is governed by the TEV's superheat setting. This is to say that even if a systems evaporator coil and TEV orifice were increased in size to the extent that the evaporator could satisfy the compressors capacity rating with a TD lower than the TEV's superheat setting, the TEV would still throttle refrigerant flow to affect its superheat setting.

Consider that it's the difference in temperature between vapour leaving the evaporator and saturated liquid near the end of the evaporator that constitutes superheat. Also, that the highest temperature the leaving vapour could warm to depends on the warmest medium it comes into contact with before leaving the evaporator, which is usually assumed to be the returning cooled medium.

While, ordinarily, heat-exchangers are configured counter-flow in order that log mean temperature difference (LMTD) be maximized for most effective use of the heat-exchangers surface area, direct expansion or dry expansion (DX) evaporators are configured for counter-flow for a slightly different reason which is simply to maximize the available TEV superheat. This means that the evaporator dry line, or otherwise suction line, would be facing, or in contact with, the warmer returning cooled medium as opposed to the colder leaving cooled medium.

Perhaps the standard TEV superheat setting for standard water chilling evaporators is around the 4K mark. If we consider that often water chilling systems are, either designed to or commissioned to, run with full load return and supply water temperatures of 10°C and 5°C over a saturated evaporating temperature of 2°C, it is clear then that with a counter-flow configuration the maximum available design superheat would be 8K while with a parallel-flow configuration the maximum available design superheat would be a mere 3K which is 1K below our assumed TEV superheat setting of 4K.

A parallel flow configuration wouldn't necessarily have any substantial effect on the evaporators LMTD because on the refrigerant side of the heat-exchanger we have a near constant temperature phase-change fluid. The TEV however, at full load, would be forced to throttle saturated suction pressure down to a temperature 4K below supply or leaving water temperature, this in order to affect the TEV's fully loaded 4K superheat setting, meaning a saturated suction temperature of 1°C (5°C-4K) would prevail. Further, if the water chiller's leaving temperature control is programmed to maintain a dead-band of 3K, meaning water leaving temperatures would fluctuate 1.5K either side of set-point, then the water chillers leaving temperature would be dropping to 3.5°C (5°C-1.5K) before a compressor capacity stage is removed which would permit a temporary saturated suction temperature of -0.5°C before unloading. This can be expected to occur any time the compressor unloads from a fully loaded sate to one stage below fully loaded, which then brings about a welcome saturation suction pressure rise and too a TEV superheat reduction. The TEV superheat reduction comes about as the TEV rides down it's Capacity/Superheat curve. Of course, if the water chiller LP setting is at an equivalent to a saturated suction temperature of 0°C, as it should be to avoid a freeze up, then suddenly the likelihood of an LP trip becomes very real.

The problem with the idea of reducing the TEV's superheat setting to avoid experiencing sub zero saturated suction temperatures is that the valve will likely hunt, especially at low loads, continually over and underfeeding the evaporator, affecting reduced evaporator efficiency and hence reduced system efficiency. Liquid flood backs are of course also a common result of low TEV superheat settings and hunting, which would very likely mean that the frequent LP trips would simply be replaced instead by frequent compressor oil failure trips caused by the compressor oil pump being fed a foaming refrigerant diluted oil. The commissioning engineers had decided in the end that to raise saturated suction pressures sufficient to avoid the frequent nuisance LP trips they would raise the water chillers leaving temperature set point by 2K from 5°C to 7°C. Although this did solve the LP problems it seemed to have increased the conditioned space's high humidity problems.

The cause of increased space humidity's.

Originally, extra to the systems frequent LP trips, the conditioned space's humidity's were occasionally but repeatedly spiking causing nuisance high relative humidity (RH) alarms. This we guessed was simply the readmitting of coil condensate back into the supply air as the chilled water temperature increased to a temperature toward return air dew point every time the compressors tripped on LP. But the high RH conditions became a more permanent problem immediately after the water chiller's nuisance LP trips had been taken care of. The newer high RH problems were of course a result of the increased AHU's coil temperatures brought about by the increased water chiller's set point.

An air conditioning system's total heat load consists of two distinctly different loads, sensible and latent. The first is merely the cooling of the air, the second is both the cooling and the condensing of moisture vapour out from within the air. Sensible heat has to be removed from the conditioned space at the same rate it is being introduced to the conditioned space and the same is true of latent heat. The ratio of the sensible heat to the total heat, sensible + latent, is called the Sensible Heat Ratio (SHR).

For any given return air dry and wet bulb temperature the AHU's SHR is dependant on effective coil temperature, a temperature known to some as the Apparatus Dew Point (ADP). ADP is always a few degree's above the temperature of the cooling fluid within the coil tubes per the temperature gradient that need exist between that fluid and the cooled medium. ADP is normally about 3K warmer than the coils contained cooling medium. Once the coil's ADP has dropped below return air dew point (DP) then for any further drop in ADP the coils sensible heat ratio reduces while the total coil capacity increases. Total capacity increases due to increased coil to return air TD. For a constant ADP an AHU's SHR will increase with increased dry bulb temperatures and decrease with increased wet bulb temperatures. Or for a constant dry and wet bulb temperature, an AHU's SHR would decrease with reduced ADP's. With DX AHU's, reduced ADP's can be easily achieved by reducing air flow which in turn reduces air to coil LMTD tending to drop saturated evaporating temperatures affecting a decrease in SHR for improved moisture removal. But reducing the air flow over a chilled water fed fan coil doesn't reduce feed water temperatures, they're controlled by the water chillers set point and control dead band, only leaving water temperatures are reduced which has little effect but for a loss in system capacity especially since the fed water temperatures are already too high.

By my experience, other common causes of increased space humidity's, or the pre-required increased saturated evaporating temperatures, have been the likes of increased head pressures, operating parameters altered by refrigerant retrofits or lost compressor efficiencies from wear and tear. Increased head pressures causing increased ADP's and hence increased space humidity's can usually be associated with fouled condensers, fouled condenser fan blades or slipping condenser fan belts. With retrofits or drop-in replacements, where say the TEV's capacity has been increased due to increased pressure drops and increased net refrigeration effect coinciding with lost compressor volumetric efficiencies due to increased compression ratios, ADP's are again caused to rise. If overall system capacity has been increased by the drop-in then subsequent reductions in compressor run time will further reduce overall system SHR.

The solution recommended in this particular case was to re-connect the chilled water piping to the return and supply headers the correct way round.

Marc O'Brien can be contacted via the Fridgetech.Com Ltd website at http://fridgetech.com or by mobile phone on 07734 193858.