Introduction

In the recent years, population growth, economic development, the emergence of new industrialized countries and also the increase in the human’s life quality have caused significant growth of energy consumption in the world like taking 10–20% of total electricity of the world by air conditioning systems [1]. The overall energy consumption disadvantages include non-renewable energy resource depletion, environmental pollution caused by emission of carbon monoxide (CO), changing weather patterns across the world, and reduction in the ozone layer thickness [2]. These factors have led many ways of renewable energy for cooling to be investigated widely. One of these solutions which is very reliable is the use of solar energy to produce part of the driving force required for cooling system whether compression or absorption. In the absorption cycles, the sun’s energy is used to provide the required heat in generators; this heat content in single effect chillers is about 70–90 °C [3]. On the other hand, when the air conditioning system has to do dehumidification and reduce temperature simultaneously, the latent and sensible loads need to be separated so that they can be controlled effectively and precisely. In order to control and isolate latent load from sensible load in 1955, the desiccant cooling systems were introduced. Desiccant cooling systems through absorbing moisture in the desiccant material are able to separate latent and sensible loads and thus are promising to save energy and reduce emissions of greenhouse gases [4]. Through hybrid systems approach, many cycles can be introduced as a substitute for air conditioning purposes. One of these units is the solar desiccant absorption refrigeration cycle which employs solar energy to supply all or part of the required heat for chiller generator and desiccant regeneration section. Solar energy becomes popular these days [5,6,7,8,9,10,11,12,13,14,15,16,17,18]. Qin et al. [19] theoretically offers a new solution to balance the heat absorption of a built dam, a solution that is perhaps the most effective way to remedy the thermal asymmetry of a built-up dam.

Given the importance of this issue, many scientists today are researching and developing solar cooling systems, solar absorption cooling, desiccant cooling, and a combination of both of these. Fung et al. [20] examined five different cooling units for Hong Kong that included: solar electric condensing refrigeration, water absorption cooling, solar absorbing cooling (solid), mechanical refrigeration compression, and solid desiccant cooling, and concluded that the and solar absorption refrigeration systems and solar electric compression had the most energy savings. In terms of economic value of solar systems, it can also be expressed that the use of these systems can be quite economical by 2030 due to rising energy prices and a significant reduction in the production cost of solar collectors. In this field, Dieter Boyer et al. [21] to produce cooling demand with minimal cost and environmental impact have carried out an economic study and concluded that given the current prices of energy and non-financial support of governments, using the solar energy system has not been economical. In relation to the desiccant technology, Henning et al. [22] came to the conclusion that desiccant unit compared to conventional systems can contribute up to 50% energy savings. Moreover, such systems’ operating costs are much lower and also more friendly with the environment. Daewoo et al. [23] demonstrated the viability of using desiccant cooling systems in various weather conditions was demonstrated and its benefits were significant in cost savings. Krishna and Mortti [24] conducted a test on the desiccant wheel containing Silica Gel and concluded with ambient air considered as process air left much impact on system performance. Vineyard et al. [25] compared desiccant system for harsh weather conditions and come to the conclusion that as the latent load share was greater in the total load, the desiccant system performance would be better. Performance of heat exchangers tried to be improved via researchers [26,27,28,29,30,31,32,33,34,35,36,37,38,39,40,41,42,43,44]. The purpose of current article was to calculate the technical and economic impact of adding desiccant wheel dehumidifier system to the common absorption cooling system. Due to the high humidity level of this project site, that is expected to have a positive effect. However, regarding extremely high radiation, solar hot water was utilized to supply hot water of chiller.

The purpose of this paper is to technically and economically examine the solar desiccant cooling system. Since typically the solar systems efficiency is not high, and using these systems for many functions and areas cannot be economical, therefore, as an innovation in this study and with the purpose of maximum efficiency of solar energy, to provide the required heat for desiccant wheel regeneration of hot H2O returned from the chiller with a temperature of about 83 °C, which is primarily used to heat the regeneration air and then poured into the hot water storage tank and the collector. This will not only reduce support fuel consumption compared to the conventional hybrid cycles but also makes temperature of the water entering the collector drop; and this is expected to augment the efficiency of solar collector units.

Description systems

The sample that has been examined in this study is a single-story computer building with an area of 400 square meters in the city of Bandar Abbas which contains 150 people in peak condition. A weather condition at the project site is shown in Table 1.

Table 1 Weather information of the site
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Based on the above data and by HAP4.5 software, maximum cooling load in summer is 40 tons of refrigeration, total air discharge required is 16000cfm, and sensible heat ratio (SHR) is 0.473 for the building.

For the mentioned building cooling load, three hybrid cooling systems were considered and modeled in the TRNSYS software. The three systems were: solar absorption cooling cycle as basis cycle, solid desiccant cooling cycle combined with solar absorption cooling and direct air of regeneration, and solar desiccant absorption cooling cycle using heat returning of chiller generator for the regeneration as the original cycle that in the following each cycle components and method of operation will be discussed.

Solar absorption cooling cycles (basis cycle)

The cycle scheme is given in Fig. 1. As shown below, this cycle is composed of three main components: 1) the hot water supply required for chiller, 2) the cold water supply for cooling coils, and 3) the supply of interior comfort conditions by the air handling units.

Fig. 1
figure 1

Solar absorption cycle scheme

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Thus, each section contains a subset of components to meet the expected output of the sector. To get the required heat in the chiller generator, solar flat plate collectors have been applied. In order to ensure the required water temperature at all hours of the day, before a tank with auxiliary heater is used before chiller generator incoming. To provide the required cooling load in the cooling air, an air handling unit is used which in a cold water supply of a cooling coil is provided by an absorption chiller which is based on the load on the chiller and temperature of the returned water from coil into it deals with cold water supply with 7 °C temperature through heat absorption in generator.

Solid desiccant cooling cycle combined with solar absorption cooling and direct heating of air regeneration

In this sector due to high latent load and low sensible heat ratio, in order to save energy and increase thermal performance of the cooling cycle’s equipments, a set of dehumidifiers like desiccant wheel are added to the previous cycle which contains three main parts of desiccant wheel, heat wheel, and direct heater of the returned air. Providing latent load supply is carried out by desiccant wheel with silica gel, and the schema of the cycle is shown in Fig. 2.

Fig. 2
figure 2

Schema of the solid desiccant cooling cycle combined with solar absorption cooling and direct heating of air regeneration

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For continuous operation of desiccant wheel and doing dehumidification, in the desiccant material in dehumidification wheel are required to be retrieved after the stage of dehumidification, restore the moisture to the regeneration air and gets ready for next dehumidification. For the proper conduct of this process, the regeneration air must be heated to the proper level that as shown in Fig. 2, this heat is applied directly to the returned air by gas air heater.

Solar cooling absorption desiccant cycle using return heat of the chiller generator for the regeneration sector

The solar systems often have a low efficiency and are not cost-effective; one of the parameters affecting the efficiency of collector is the fluid temperature entering into it. On the other hand, temperature of the hot water returning from chiller generator is about 80–83 °C that is greater than the needed temperature in this project is to restore the desiccant wheel. So as a useful and innovative solution, the output hot water from the chiller generator, as shown in Fig. 3, first needs to be entered in a water-to-air exchanger before the desiccant wheel and thereby the restoring water temperature gets to the required level. Then in the following, this water goes into the storage tank and from there goes into the collector with a temperature less than that in the previous cycle. After doing this, the external energy consumption is reduced and the collector efficiency and solar system are increased.

Fig. 3
figure 3

Schema of the solar desiccant absorption cooling cycle using chiller generators return heat for regeneration

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Systems modeling

To check the performance of each system and the influence of various parameters on the output conditions, the three above cycles are modeled based on Figs. 1–3 in the TRNSYS software. Accordingly, in Fig. 4, the third version of the system (final system) can be seen in this application. In the above diagram, the type686 acts like a hypothetical building, and we enter the output of the carrier software in this type. Then, the output latent load of the type686 will be entered in type693. In the following, 60% of the output air of 693 and 40% of the total inflow from the ambient air enter into type648, and this type, based on the psychometric principles, mixes the two inputs and calculates the output air condition. Then the outlet air from the type648 enters into type506d, which has the function as air conditioner. The output air of type506d enters into type696, and it calculates the amount of transient heat. From the outlet air of type696, two outputs are taken: one goes to type693 to complete the ventilation cycle, and the other enters the chiller’s size to 682. The flow of fluid from 682 in the return path as cold water enters the chiller, or the same type107. China’s hot water is also provided by type537. The weather conditions of the site are simulated by type109 and entered into the input of type537, and then the heated water is pumped into 537 pumps of 695 and then the hot water enters the kettle reservoir. Typing 13 acts as a pressure relief valve in this direction and determines the amount of energy that needs to be repelled so that the temperature is always lower than the boiling time. The tank is four layers of reservoir type, and the water first enters the highest layer. After leaving the storage tank, the water enters a gas-fired furnace called type659, where it is heated up to reach the needed inlet T = 107. In this diagram, the type683 is the same as the slow-moving handwheel and the inlet air before it enters the type706. Typing 91 is also used as a desiccant wheel reclaimer that obtains the required heat from the chiller’s hot water outlet heat.

Fig. 4
figure 4

The modeled plan of the third cycle in TRNSYS software

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The return water temperature from cooling coil of air handling unit (type696) was obtained according to Eq. 1:

$$ T_{\text{chw,out}} = \frac{{\dot{Q}_{\text{rem}} }}{{\dot{m} \cdot C_{\text{Pchw}} }}T_{\text{chw,set}} $$
(1)

Then a fraction of the time of that chiller (type107) needs to function and actual capacity of chiller was obtained based on the relations 2 and 3:

$$ f_{\text{designload}} = \frac{{\dot{Q}_{\text{rem}} }}{{{\text{Capacity}}_{\text{Rated}} }} $$
(2)
$$ {\text{Capacity}} = {\text{f}}_{\text{FullLoadCapacity}} \times {\text{f}}_{\text{NominalCapacity}} \times {\text{Capacity}}_{\text{Rated}} $$
(3)

In the above equations, CapacityRated is given as input by the user. Then the output cold water of the chiller enters into the cooling coil of the air handling units that have a set point of output temperature of 18 °C. In this device according to 4 and 5 equations, the latent and sensible heat of the air must be used to achieve the best output conditions.

$$ \dot{Q} = \dot{m} \times (h_{\text{in}} - h_{\text{out}} ) $$
(4)
$$ \dot{Q}_{\text{Lat}} = \dot{m} \times (\omega_{\text{in}} - \omega_{\text{out}} ) $$
(5)

In the cycle’s thermal driver part, solar energy and flat panel collector (type537) are used to provide the needed hot water of the chiller due to the high radiation intensity in the project site (annual average over 22000 kWh m−2). However, there is a hot water tank auxiliary and heated of gas as a backup before entering the water into chiller generator to ensure that hot water is available 24 h a day and the temperature of the chiller input water will be 88 degrees Celsius all the time.

Collector used in the present article has a controlled pump in its output, so that the pump with changing of the output discharge always provides the temperature of 88 °C in the output, and as it is notable to provide the temperature, the discharge gets zero.

Accordingly, the collector is evaluated using Eq. 6 of the collector useful heat absorption and then Eq. 7 of fluid discharge flowing out of the collector:

$$ \dot{Q}_{\text{u}} = A_{\text{c}} [F_{\text{R}} (\tau \alpha )I_{\text{t}} - F_{\text{R}} U_{\text{L}} (T_{\text{i}} - T_{\text{amb}}^{{}} )] $$
(6)
$$ \dot{m}_{\text{fluid}} = \left( {\frac{{\dot{Q}}}{{C_{\text{p}} (T_{\text{i}} - T_{\text{out}} )}}} \right) $$
(7)

That for the collector used values of FR (τα) and FR UL are, respectively, 0.8 and 5.0. Due to high humidity in the project site and high absolute humidity the entering into the cooling coil (0.015 kg water/kg air) and output humidity set point at 0.008 kg water/kg air, and since air handling units have been required to provide both sensible and latent loads by themselves, the energy consumption is expected to be high in this mode.

Through adding a desiccant wheel set as depicted in Fig. 4 (type653), the output air first enters into the desiccant wheel and there its absolute humidity is reduced down to 0.008. Dehumidification is done by the desiccant wheel as shown in Fig. 5 and by the F1 and F2 co-potential lines through 8 and 9 equations.

$$ F1 = \frac{ - 2865}{{T^{1.490} }} + 4.344\omega $$
(8)
$$ F2 = T1.490 - 1.127\omega 0.07969 $$
(9)

In the following process, a heat wheel exchanger (type760b) is posited for the purpose of heat and energy regeneration from returning air inside the building, that in fact has been an eddy air exchanger and itself has led to more decrease in sensible heat.

Fig. 5
figure 5

F1 and F2 co-potential lines for silica gel

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Minimum heat capacity, enthalpy of the air at the outlet of this device and the sensible heat content between exchanged two streams are obtained, respectively, through 10–12.

$$ C_{\hbox{min} } = \hbox{min} (\dot{m}_{\text{exh}} C_{\text{Pexh}} ,\dot{m}_{\text{fresh}} C_{\text{Pfresh}} ) $$
(10)
$$ \dot{Q}_{\text{sens}} = \varepsilon_{\text{sens}} C_{\hbox{min} } (T_{\text{exh,in}} - T_{\text{fresh,in}} ) $$
(11)
$$ h_{\text{fresh,out}} = h_{\text{fresh,in}} + \frac{{\dot{Q}_{\text{sens}} }}{{\dot{m}_{\text{fresh}} }} $$
(12)

On the back track as well, the air after flowing within the heat exchanger to reach the required temperature for recovery desiccant wheel and as the result of doing continuous and accurate dehumidification measurement may pass through the two paths or, according to Fig. 3, enters directly into gas heaters and then into the desiccant wheel on the circumstances according to Eq. 13.

$$ h_{\text{air,out}} = h_{\text{air,in}} + \frac{{q_{\text{in}} }}{{\dot{m}}} - \frac{UA}{{\dot{m}}}(\bar{T} - T_{\text{amb}} ) $$
(13)

Or according Fig. 4 and the optimized cycle enters into a water-to-air exchanger (type91) and is heated by the returning hot water from the chiller generator, and then enters into the desiccant wheel. By this, the direct heater for the regeneration air is no more required and this causes to save energy and cost.

Economic analysis

The economic analysis of solar systems uses a variety of methods and criteria which includes Life Cycle Cost (LCC), Life Cycle Saving (LCS), Payback Time (PT) and Return on Investment (ROI).

In this paper, LCC method is used to estimate the cost of each system. According to this method, the total current and fixed costs of the system can be calculated during the period of operation. The initial cost of the solar system is obtained through relation 14.

$$ Cs = C_{\text{A}} A_{\text{G}} + C_{\text{v}} V_{\text{s}} + C_{\text{E}} $$
(14)

where \( C_{\text{E}} \) is equal to 12% of collector and tank costs. The fuel current cost per 1 month is calculated through relation 15 and per year through relation 16.

$$ C_{\text{aux}} = {\text{day}} \times \frac{{Q_{\text{aux}} }}{{L_{\text{v}} }} \times C_{\text{g}} $$
(15)
$$ C_{\text{auxY}} = \sum\limits_{i = 1}^{12} {C_{\text{aux}} } $$
(16)

The point that should be noted here is that the current prices were paid during the years of unit operation, while the initial cost is paid in the first year. Therefore, different year’s costs ought to be calculated in the rate of the first year so that they can be sum up. For this purpose, a factor called the present worth factor is used which can be calculated using Eq. 17:

$$ {\text{PWF}}(N,I,d) = \sum\limits_{j = 1}^{N} {\frac{{(1 + i)^{j - 1} }}{{(1 + d)^{j} }}} = \left\{ {_{{\frac{N}{1 + i}}}^{{\frac{1}{d - i}\left[ {1 - (\frac{1 + i}{1 + d})N} \right]}} \mathop{\longrightarrow}\limits{{}}} \right._{{_{{{\text{if}}(i = d)}} }}^{{_{{{\text{if}}\left( {i \ne d} \right)}} }} $$
(17)

In the above equation, N is number of year’s number of the system lifetime, i is inflation rate, and d is the discount rate. On this basis and in accordance with relations 18 and 19, the total cost of each system is calculated separately in the rate of the first year and the superior cycle introduced from the economic perspective.

$$ C_{\text{auxT}} = C_{\text{auY}} \times {\text{PWF}}(N,i,d) $$
(18)
$$ {\text{LCC}} = C_{\text{s}} + C_{\text{auxT}} $$
(19)

In the economic analysis of systems in the present study, the following hypothesis has been considered:

  • The market discount rate equal to 20%.

  • The market inflation rate equal to 20% and fuel price inflation rate equal to 15%.

  • Number of years for system performance equal to 20.

  • The whole system at the end of a 20-year period was sold to the rate of 50% of its initial price.

  • The entire solar sector of each system including collector and storage tank are provided with loans funded for 15% interest and 20-year repay.

  • The cost of the equipment electricity consumption is neglected.

  • The efficiency of the whole cycle during the functioning year does not change.

Results

It was noted in the previous section that the above modeling has been developed for a computer building in Bandar Abbas, Iran, with 400 meters of area and population occupation. Type of the used collector was also flat with an area of 300 square meters and the storage tank was of multi-node type to the total volume of 12 m3. According to the modeling system in TRNSYS and analysis of the output parameters, results of this modeling are given in Table 2.

Table 2 Results of the cycles modeling
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In estimating the system gas consumption prices, the sales price per 1 m3 of natural gas is considered to be equal to 0.28 $ based on the general policy of the state and due to the gradual elimination of energy subsidies. Therefore, for clear understanding of the issue, the economic analysis once is done with the current price of natural gas (0.016 $ per cubic meter) and another time on the basis of the intended final price.

Figures 6 and 7 also, in a more clear way, show comparison of three cycles. In this diagram, the parameters of basic cycle are considered 100%, and other related parameters of the 2 other cycles have been measured in percentage.

Fig. 6
figure 6

Comparison between the basic parameters in cycles

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Fig. 7
figure 7

Comparison between the basic parameters in cycles (solar fraction)

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As can be seen in Fig. 6, the cooling load imposed on the third cycle as the result of the optimization cycle dropped to 42.2% compared to the base cycle and to 1% compared to the second cycle. Also, the energy content needed in auxiliary heaters, as expected, in the third cycle dropped to 29% compared to the base cycle and to 4.6% compared to the second cycle. On the other hand, as in the basic and second cycles the water temperature entering into the collector is the same, so the collector useful heat gain in these two cycles is the fixed value of 750,000,000 kJ, while considering drop of this temperature in the third cycle, this parameter is increased to 8% compared to the two previous cycles which is a significant amount.

After cooling load reduction of the cycles 2 and 3, the chiller absorption capacity used in the third cycle dropped to 31% compared to the base cycle and to 4.8% compared to the second cycle. However, regarding the reduction in energy consumption by auxiliary gas heaters, it has been observed that the natural gas consumption over 20 years of system functioning in the third cycle dropped to the rate of 29% compared to the basic cycle and to the rate of 4.6% compared to the second cycle that is very desirable due to the limited energy resources and the environmental impact.

Finally, after economic analysis of the three cycles based on costs method over the lifetime and according to Fig. 6, it can be seen, including the price of gas as 0.016 $ per cubic meter, the total cost of the third cycle dropped to the rate of 3.3% compared to the basic cycle and 1.7% compared to the second cycle. However, with the final intended price for natural gas, namely 0.28 $ per cubic meter, the third cycle total cost dropped to the rate of 18% (equivalent to 6.58 billion Rials) compared to the first cycle and 4% (equivalent to 35,695 $) compared to the second cycle reduced.

Conclusions

According to the diagram in Fig. 6 and Table 2, parameters related to energy consumption, fuel, and costs have taken a decreasing trend from the first to the third cycle, which is desirable. The uptake of collector useful heat has a direct impact on collector efficiency and the solar part of the cycles. In cycles 1 and 2, due to the equal collector inlet water temperature, the value is constant, but in the third cycle due to the use of water returning from the chiller desiccant wheel in the regeneration process, this value increased which is also useful. Consequently, the third cycle with 1,371,980 m3 of natural gas savings and 156,667 $ cost savings compared to basis cycle was selected as the most optimal cycle, both environmentally and economically.